Htdrodynamic bearing assembly and spindle motor having the same

ABSTRACT

This invention is to provide a hydrodynamic bearing assembly, which realizes the high rotation rate in a stable manner and the robust rigidity. The hydrodynamic bearing assembly has a total radial gap of 3 microns or less for preventing the contact in the thrust bearing. The thrust bearing is a pomp-out type one, and the radial bearing has offset grooves on the surface thereof to supply the fluid flow sufficiently to the thrust bearing. The grooves also eliminate a half-whirl. A depth ratio relative to the diameter of bearing is preferably 0.005 or less to avoid the reduced translation rigidity. The radial gap is smoothly enlarged from the center to both ends along the axis, and the shaft is biased to incline the shaft relative to the sleeve, so that the shaft can be rotated with the robust rigidity. Also, a pair of the thrust bearings is provided on both ends of the radial bearing so as to realize the robust rigidity.

TECHNICAL FIELD

This invention relates to a hydrodynamic bearing assembly, and inparticular, relates to the hydrodynamic bearing assembly incorporatedwith a spindle motor used for driving a memory device such as a harddisk drive (referred to as a “HDD”, hereinafter), or bar code reader.This invention also relates to the spindle motor including thehydrodynamic bearing assembly, as well as the memory device and the barcode reader including the spindle motor.

BACKGROUND ART

The conventional spindle motors used for driving the memory device suchas the HDD or a bar code reader, includes a hydrodynamic bearingassembly implementing high rotation in a stable manner for a longeffective lifetime. Among various features of the hydrodynamic bearingassembly, since rotational member and stationary member of thehydrodynamic bearing assembly operate without contact to each other, therotation thereof causes mechanical friction therebetween less than thatof another contacting type of the bearing assemblies such as a ballbearing. In comparison with the hydrodynamic bearing assembly using oilfor generating dynamic pressure, in particular, the hydrodynamic bearingassembly using gas as fluid has further advantages to reduce the coomcaused by shattering lubricant such as oil and grease in addition to thereduced friction.

FIG. 42 shows an exemplary spindle motor with a conventionalhydrodynamic bearing assembly. In the drawing, the hydrodynamic bearingassembly comprises, on a base plate 200, a column shaft 201, a sleeve202 rotatably arranged around the shaft 201 leaving a predetermined gapalong the axis direction of the shaft 201 for relative rotationtherebetween. The hydrodynamic bearing assembly also comprises a thrustplate 202, which is arranged perpendicular to the shaft 201 and opposesto the bottom surface of the sleeve 202. A radial bearing is formedbetween an outer surface of the shaft 201 and the inner surface of thesleeve 202. Also, a thrust bearing is formed between the bottom surfaceof the sleeve 202 and the thrust plate 203. The thrust plate 203includes grooves 205 for generating thrust dynamic pressure, formed onthe surface opposing to the bottom surface of the sleeve 202, asillustrated by a dashed line.

In this specification, the bottom surface opposing to the thrust plate203 and defining the thrust bearing in cooperation therewith is referredto as a thrust opposing surface. In FIG. 42, one of the end surfaces inthe axis direction is the thrust opposing surface 204. A rotor 207attached with the sleeve 202 can be rotated about the shaft 201 with thesleeve 202. The rotor 207 has a rotor magnet 208 arranged on the innersurface of a skirt 207 a of the rotor 207. The rotor magnet 208 opposesto the electromagnet 209 arranged on the base plate 200. In case of theHDD, a plurality of memory media are mounted on the outer surface, alsoin case of the bar code reader, a polygonal mirror is mounted on therotor 207, both of which rotate with the rotor 207.

According to the spindle motor constructed as described above, analternating current supplied to the electromagnet 209 causes theattraction and/or repulsion forces between the electromagnet 209 and therotor magnet 208. This provides the rotor 207 supporting the rotormagnet 208 with a rotation drive force so that the rotor 207 and thesleeve 202 attached therewith together rotate around the axis of theshaft 201. The rotation causes the relative movement between the shaft201 and sleeve 202, generating the radial dynamic pressure due to thefluid in the radial bearing. In general, although air is often used forthe fluid intervening between the shaft 201 and sleeve 202 when thespindle motor is used in the atmosphere, particular gas or oil may beused as the fluid. In this specification, the intervening object forgenerating the dynamic pressure is referred to as the “fluid”. Theaforementioned rotation also causes the relative movement between thethrust opposing surface 204 of the sleeve 202 and the thrust plate 203,thereby generating another dynamic pressure in a thrust direction due tothe grooves 205. To this end, this thrust dynamic pressure allows therotational member such as sleeve 202 and rotor 207 to rotate about theshaft 201 keeping the rotational member away from the stationary membersuch as shaft 201 and the base 200.

FIG. 43 shows the thrust grooves 205 formed on the surface of the thrustplate 203 for generating the thrust dynamic pressure in the thrustbearing. As shown, the grooves 205 include a plurality of a spiralgroove, each of which is angled at a predetermined angle with the circleon the thrust plate 203, and has a depth in a range of 1 micron through10 microns. The thrust opposing surface 204 of the sleeve 202 rotates ina direction indicated by the arrow 206 against the grooves 205 so thatthe fluid such as air is convolved in the grooves 205. The fluid ispressed along the spiral grooves 205 towards the axis due to theviscosity of the fluid during the above-mentioned rotation, hereby togenerate the pressure (dynamic pressure). This dynamic pressure operatesthe thrust opposing surface 204 to push up the rotational member such assleeve 202. Such bearing assembly, which conducts the fluid from thecircumference towards the axis of the thrust bearing assembly togenerate the dynamic pressure, is referred to as a “pump-in” bearingassembly. The pump-in bearing assembly is commonly used for thehydrodynamic bearing assembly.

A need has been existed in the market to a compact and lightweighthydrodynamic bearing assembly implementing the rotation at high rate andheavy load in a stable manner. There are some problems to be solved forthe hydrodynamic bearing assembly to satisfy such market's needs.Firstly, the rotation should be stable in particular at the highrotation rate. Secondary, the bearing assembly should have a certainrigidity sufficient to bear against the oscillation forces provided fromexternal circumstances. Thirdly, the bearing assembly has to be improvedin the activation feature to activate rotation of the rotational memberin contact with the stationary member. Fourthly, the bearing assemblyshould be more compact and lightweight. Details for those problems to besolved will be described hereinafter.

(First Problem)

In order to address the first problem, i.e., to realize the highrotation rate in a stable manner, it is necessary to eliminate aphenomenon, so-called half-whirl. The half-whirl is the phenomenonappeared due to the rotation of sleeve 202 relative to the shaft 201with a predetermined gap for keeping thereof away to each other. Thefluid intervening between the outer surface of the shaft 201 and theinner surface of the sleeve 202 for generating the dynamic pressurecauses a continuous pressure distribution therebetween due to therelative rotation. When the external disturbance causes either one ofthe shaft 201 or sleeve 202 to deflect from the rotation axis, the forcedue to the dynamic pressure is offset to the rotation axis so that thehorizontal component of the force revolves the rotational member aroundthe rotation axis without returning the rotational member to itsoriginal position. The convergence of the revolution returns therotational member to the original position so that the rotational memberrotates in a stable manner. On the contrary, if the revolution is kept,the rotational member whirls around the central axis of the stationarymember resulting in the unstable rotation. This phenomenon is referredto as the half-whirl. The present inventors have discovered that therevolution is likely to be kept with the bearing assembly having thecontinuous pressure distribution in comparison with one having adiscontinuous pressure distribution.

FIG. 44 schematically illustrates the half-whirl phenomenon, showing across section along the rotation axis of the stationary shaft 201 andthe rotating sleeve 202 of the hydrodynamic bearing assembly. In thenormal operation, the sleeve 202 rotates around a rotation centerconcentric with the stationary axis I of the shaft 201 as indicated bythe cross (+) in the direction of the arrow 215. When the externaldisturbance causes the sleeve 202 to deflect relatively to the shaft201, the rotation center of the sleeve 202 of the rotational member isshifted from the stationary axis I to the position C as indicated by thealphabet (X). The force generated by the dynamic pressure having thedeflecting direction as well as the continuous pressure distributionrotates the sleeve 202 on its own axis, and also revolves the rotationaxis C of the sleeve 202 around the stationary axis I along the arrow216 in a whirling manner. For example, the dashed line illustrates thesleeve 202 a after the rotation axis C of the sleeve 202 indicated bythe solid line revolves 180 degrees around the stationary axis I. Inthis instance, the rotation axis C of the sleeve 202 is shifted alongthe arrow 216 to the rotation axis C′. The half-whirl whirls therotational member (such as sleeve 202 in FIG. 44) relatively to thestationary member (such as shaft 201 in FIG. 44) so that the bearingassembly loses the stability in rotation, thereby to cause undesiredoscillation and/or malfunction of the bearing assembly used for the HDDor bar code reader.

(Second Problem)

The second problem to be solved, i.e., the rigidity/stiffness of thehydrodynamic bearing assembly will be discussed with reference of FIG.45. This drawing is the enlarged view of the hydrodynamic bearingassembly of FIG. 42, in which similar reference numerals denote thesimilar components. In the drawing, the parallel lines schematicallyillustrate the dynamic pressure distribution generated during therotation of the hydrodynamic bearing assembly. The dynamic pressuredistribution M is generated in the radial bearing defined between theshaft 201 and the sleeve 202 m, thereby to keep them away from eachother. On the other hand, the dynamic pressure distribution N is alsogenerated in the thrusts bearing defined between the thrust plate 203and thrust opposing surface 204 so that no contact is kept therebetween,allowing the sleeve 202 of the rotational member to rotate without anycontact.

The hydrodynamic bearing assembly of FIG. 45 receives external forcesincluding a force indicated by the arrow 217 perpendicular to thebearing axis (translation force), a force indicated by the arrow 218along the bearing axis (elevation force), a force indicated by the arrow219 around an axis perpendicular to the bearing axis (oscillationforce), and the combination thereof. The hydrodynamic bearing assemblyis required to have a rigidity against such forces enough to keep therotational member away from the stationary member and to ensure thestable rotation.

FIG. 46 provides an example where the sleeve 202 is inclinedcounterclockwise relative to the shaft 201 and the thrust plate 203because of the disturbance (external forces) to the hydrodynamic bearingassembly during the stable rotation as shown in FIG. 45. In thisinstance, the shaft 201 moves closer to the sleeve at the right-upperportion indicated by T and at the left-lower portion indicated by U,also the thrust plate 203 moves closer to the thrust opposing surface204 at the leftmost portion indicated by V. In general, the wedge effectdue to the convolution of the fluid between relatively moving bearingmembers is increased as the gap therebetwen is decreased. Thus, thedynamic pressure distribution is shifted from as illustrated in FIG. 45to that as shown in FIG. 46. The dynamic pressure is increased betweenthe rotational member and the stationary member at the portions T and Uso that the repulsion force is generated to prevent both members frommoving closer to each other. The contact between the shaft 201 and thesleeve 202 is avoided unless the disturbance force overcomes therepulsion force.

Meanwhile, the fluid is guided from the circumference of thrust plate203 towards the axis (the pump-in bearing assembly) so that the dynamicpressure between the thrust plate 203 and the thrust opposing surface204 is increased towards the bearing axis, as shown by the portion N inFIG. 45. Thus, the peak of the dynamic pressure can disadvantageously beexpected at the radially outer portion V shown in FIG. 46, even if therotational member moves closer to the stationary member. Therefore, whenthe disturbance force causes oscillation force, the thrust plate 203 islikely to physically contact with the thrust opposing surface 204. Oncethe thrust plate 203 contacts with the thrust plate 203, the frictionforce therebetween results the unstable rotation of the rotationmembers. Further, the rebound followed by the contact causes theundesired impact, which could bring the malfunction of the magnetic headused for the HDD, or result an extensive damage to the spindle motor.Therefore, it is particularly important that the hydrodynamic bearingassembly has sufficient rigidity against the oscillation force, which isreferred to as “tilt rigidity”. Also, the rigidity against thetranslation force and the rigidity against the elevation force, whichare referred to as the “translation rigidity” and the “elevationrigidity”, respectively. Both of the translation rigidity and elevationrigidity can be improved by increasing the radial and thrust dynamicpressure.

FIG. 47 illustrates an exemplary hydrodynamic bearing assembly includingthe shaft 201 and the thrust plate 203 that is not perpendicularlyattached thereto. The shaft 201 is inclined relative to the thrust plate203, and the sleeve 202 is provided around the shaft 201. During therotation of the spindle motor, in general, the radial bearing and thethrust bearing of the hydrodynamic bearing assembly have gaps ofapproximately 3 to 5 microns and approximately 2 to 10 microns,respectively, between the rotational member and the stationary memberfor the rotation without contact. In order to stabilize the rotation ofthe bearing assembly, the above-mentioned gaps are kept constant in aprecise manner. In the radial bearing, since the inner surface of thesleeve 202 opposes to the outer surface of the shaft 201, the gaptherebetween can readily be kept constant. On the other hand, the gapbetween the thrust opposing surface 204 of the sleeve 202 and the thrustplate 203 in the thrust bearing is more difficult to be kept constantthan that in the radial bearing, because the gap in the thrust bearingis more susceptible to the arrangement of the sleeve 202 relative to theshaft 201. Thus, the precise control of the gap in the thrust bearingdepends upon directly how the shaft 201 is arranged perpendicularly onthe thrust plate 203 in a precise manner. Therefore, as illustrated inFIG. 47, in case where the shaft 201 is inclined to the thrust plate203, the sleeve 202 rotate about the bearing axis inclined to the thrustplate 203, even if the normal dynamic pressure is generated in theradial bearing. This inclined rotation against the thrust plate 203 mayraise a possibility that the sleeve 202 contacts with the thrust plate203 in the portion V of FIG. 47 due to a slight oscillation force duringthe rotation.

(Third Problem)

The third problem to be solved is an improvement of the actuationfeature of the hydrodynamic bearing assembly. When the hydrodynamicbearing assembly start to rotate, since no rotation generates no dynamicpressure, the sleeve 202 is in contact with the thrust plate 203, and insome cases, the shaft 201 also is in contact with the sleeve 202. Then,when the spindle motor is being actuated, the rotation at a relativelylow rate keeps those members in contact with each other. The rotationrate exceeding to a predetermined rate generates the dynamic pressureenough to ensure the rotation without any contacts. This predeterminedrotation rate is referred to as a “floating rotation rate”, hereinafter.Since the sleeve 202 rotates in contact with the thrust plate 203 beforethe floating rotation, there are problems of friction and overheatbetween the rotating and stationary member. Further, a greater drivingtorque is required to rotate the sleeve 202 in contact with the thrustplate 203. Thus, the higher floating rotation rate requires more timeand energy consumption to achieve the rotation without any contacts.Therefore, the hydrodynamic bearing assembly has been demanded such thatthe floating rotation rate is minimized to rotate the rotational memberkeeping away from the stationary member within the shortest time inorder to realize good endurance and less energy consumption foractivation of the bearing assembly.

(Fourth Problem)

The fourth problem to be solved by the present invention is to realizethe hydrodynamic bearing assembly to be more compact and lightweight.This need comes from the fact that devices such as memory deviceincorporating the hydrodynamic bearing assembly are demanded to be morecompact and lightweight. Also, the more compact and lightweight bearingassembly advantageously causes the rotation with contact between therotating and stationary member to wear less at the activation of thebearing assembly.

With respect to each of the problems to be solved as mentioned above,the prior art approaches to address the problems and the deficienciesthereof will independently be described hereinafter.

1. Half-Whirl

To address the problem of the half-whirl, the prior arts has proposed aplurality of notches provided parallel to the bearing axis on either oneof the outer surface of the shaft 201 and the inner surface of thesleeve 202, which are opposing and rotates relative to the shaft 202.FIG. 48 is a vertical cross section, and FIG. 49 is a transverse crosssection of the bearing assembly. As shown, the sleeve 202 is arrangedaround the shaft 201 for rotation of the sleeve 202 about the shaft 201.The shaft 201 is secured perpendicular onto the thrust plate 203, whichopposes to the bottom surface of the sleeve 202. The hydrodynamicbearing assembly comprises the shaft 201, the sleeve 202, and the thrustplate 203.

As shown in FIG. 48, three longitudinal grooves 221 are formed on theouter surface along the bearing axis. By providing grooves 221, thecontinuous pressure distribution generated between the rotating andstationary member of the bearing assembly is interrupted to avoid thehalf-whirl phenomenon.

In general, taking account of the dynamic balance during high rotation,the grooves 221 are provided on the surface of the stationary member,which may be either one of outer surface of the shaft 201 and the innersurface of the sleeve 202. However, similar advantages can be enjoyed byproviding the grooves on the surface of the rotational member. Theradial dynamic pressure is reduced locally, and if the grooves areformed on the stationary member, then the translation rigidity isreduced along the direction of the arranged grooves. This approach mayavoid the half-whirl but remains the disadvantage increasing thetendency to cause the rotational member in contact with the stationarymember along the direction of the arranged grooves.

Another prior art has suggested providing either one of the outersurface of the shaft 201 and the inner surface of the sleeve 202 with across section of a configuration such as the triangle (the round-apextriangle) shape instead of the circle. This changes the gap between theshaft 201 and the sleeve 202 so that the aforementioned continuouspressure distribution is interrupted. For example, Japanese PatentLaid-Open Publication 02-89807 discloses the non-circular bearingassembly. Yet, the non-circular bearing assembly also causes the dynamicpressure in the broader gaps to reduce the translation rigidity.

Further, another prior art of Japanese Patent Laid-Open Publication02-150504 discloses a plurality of longitudinal bands in the directionof the bearing axis including circumferential micro ground streaksformed on the inner surface of the sleeve 202. The ground streaks causesthe turbulent flow of the fluid in the radial bearing so that thedynamic pressure distribution leading the half-whirl is prevented. FIGS.50 and 51 are discussed in Japanese Patent Laid-Open Publication02-150504. FIG. 50 illustrates the shaft 201 surrounded by the innersurface of the sleeve 202. FIG. 51 is an enlarged view of a portion W onthe inner surface of the sleeve 202, in which the streak band 223comprises circumferential micro ground streaks 222. A plurality ofstreak bands is formed in the direction of the bearing axis with apredetermined distance to each other. This overcomes the half-whirlwhile maintaining the translation rigidity, which might be reduced bythe grooves formed on the inner surface of the sleeve 202. However,since each streak band 223 has to be scraped one by one, the task forscraping them is burdensome. In case where the inner diameter of thesleeve 202 has a small size, for example, in the order of severalmillimeters, scraping the streak bands 223 would be difficult.

2. Improvement of Rigidity of Bearing Assembly

Next, some conventional means having a main purpose for improving therigidity of the bearing assembly will be described hereinafter. JapanesePatent Laid-Open Publication 08-338960 discloses an improvement of theradial rigidity by providing a plurality of shallow longitudinal groovesalong the axis and an annular groove across the longitudinal grooves onthe outer surface of the shaft of the hydrodynamic bearing assembly usedfor an optical scan device. This structure supports the sleeve on multipoints against the radial motion so that the radial rigidity of thebearing assembly is improved. FIG. 52 is a vertical cross section of theoptical scan device incorporating the hydrodynamic bearing assembly. Inthe hydrodynamic bearing assembly, the shaft 231 is mounted on thehousing 230, and the shaft 231 is rotatably arranged around the shaft231. A flange 233 made of metal such as aluminum and brass is secured onthe outer surface. A polygonal mirror 234 for deflecting a laser beam isarranged on the top surface of the flange 233 by means of the spring235. Also, a driving magnet 236 is bonded on the bottom surface at theperimeter of the flange 233 by the adhesive. A stator 238 is provided onthe substrate 237 secured on the housing 230 so as to oppose to thedriving magnet 236.

FIG. 53 is the enlarged view of the shaft 231 alone illustrating itsdetailed aspect and a schematic pattern of the dynamic pressuredistribution (with its peaks). As shown, two of the parallel shallowgrooves 241, 242 are formed on the surface for generating the dynamicpressure. Also, the annular groove 243 is formed on the surface of theshaft 231. The annular groove 243 divides the peak of the dynamicpressure distribution into two peaks Q1, Q2. This prevents the rotatingsleeve 232 in contact with the shaft 231, thereby to avoid the damagesto each other. It is understood that since the dynamic pressuredistribution has two peaks, the bearing assembly with the annular groove243 has greater anti-moment rigidity against the external moment thanthat without the annular groove. However, the dynamic pressure at theportions adjacent to the longitudinal grooves 241, 242 formed on theouter surface of the shaft 231 is reduced to decrease the translationrigidity at those portions. Also, other factors such as dimensions andweights of the annular grooves 243 further reduces the translationrigidity.

Another conventional technique has proposed biasing the rotationalmember in the radial bearing along a particular direction. This definesthe minimum gap between the shaft and the sleeve at a predeterminedpoint where the high dynamic pressure is generated. Thus, it isunderstood that the bearing rigidity is improved because of the highdynamic pressure generated on the point where the minimum gap isdefined, and that the half-whirl can advantageously be avoided. This istrue as well for a complex radial-thrust bearing assembly, in which theradial bearing assembly and the thrust bearing assembly are continuouslyformed.

In particular, according to Japanese Patent Laid-Open Publication11-18357 and Japanese Patent Laid-Open Publication 11-55918, the rotormagnet is positioned eccentrically to the coil so that the shaft isbiased against the sleeve in the predetermined direction for the stablerotation. FIG. 54 shows one example disclosed in Japanese PatentLaid-Open Publication 11-55918. The rotor 252 is arranged around thestator 251, the rotor magnet 253 attached to the inner surface of therotor 252 is opposed to the stator 251 for driving the torque. As shownin FIG. 54, one stator 251 a has an arm shorter than those of theremaining stators 251. This causes the gap h1 between the stator 251 aand the rotor magnet 253 greater than the gap h2 between the stator 251and the rotor magnet 253. Thus, the attraction force (or the repulsionforce) is reduced in the gap h1 so that the rotor is biased against thestator in the predetermined direction. The stator 251 is securedconcentrically to the shaft, the rotor 252 is biased against the shaftin the predetermined direction. However, in order to bias the shaftagainst the sleeve with use of the method, the stator 251 is required tobe positioned concentrically to the baring axis. In practical, thealignment of the stator 251 is often impossible because of the designconstraint.

Japanese Utility Model Laid-Open Publication 55-36456 discloses thestationary permanent magnet attached to the housing so as to oppose tothe rotor magnet for tilting the rotor towards the predetermineddirection for rotation. However, this approach causes the lifetime ofthe bearing assembly shorter because the distal edge of the shaft 255contacts with the sleeve 254.

Some other prior arts use a pump-out type hydrodynamic bearing assemblyto improve the bearing rigidity. The hydrodynamic bearing assembly, inwhich the fluid is conducted from the center towards the circumference,is referred to as the pump-out type hydrodynamic bearing assembly.Briefly speaking, the pump-out type hydrodynamic bearing assembly hasspiral grooves 205 having the angle with the circle, or the rotatingdirection reversed to one indicated by the arrow 206 of FIG. 43. Sincethe pump-in type hydrodynamic bearing assembly has the dynamic pressuredistribution with the peak adjacent to the axis, it is relativelysusceptible to the external disturbance. Meanwhile, the pump-out typehydrodynamic bearing assembly has the peak of the dynamic pressuredistribution at the outermost edge of the thrust plate 203, thereby toimprove the rigidity against the disturbing motion.

FIGS. 55 and 56 illustrate one embodiment to implement the pump-out typehydrodynamic bearing assembly according to Japanese Patent Laid-OpenPublication 9-229053. In FIG. 55, the sleeve 272 is arranged around theshaft 271. Also, the shaft 271 has the thrust plate 273, which isintegrally formed and is flush with the surface perpendicular to thebearing axis. The shaft 271 and the thrust plate 273 together rotatewithin the chamber defined by the sleeve 272. The shaft 271 hasherringbone grooves 274 on the outer surface.

FIG. 56 is a top view of the shaft 271 and the thrust plate 273 of FIG.55. The thrust plate 273 includes a plurality of spiral grooves 275 forgenerating the dynamic pressure on both surfaces of the thrust plate 273(including opposite surface of the drawing). The arrow 276 shows therotational direction of the shaft 271 of the bearing assembly. Asillustrated, the spiral grooves 275 on both surfaces of the thrust plate273 are formed with the tilt so that they conduct the fluid within thebearing assembly radially from the center to the circumference. Also,the thrust plate 273 has a plurality of through-holes extending alongthe bearing axis adjacent to the shaft 271.

In FIG. 55, during the rotation of the bearing assembly, the herringbonegrooves 274 guide the fluid away from the thrust plate 273. Contrary,the spiral grooves 275 of the pump-out type hydrodynamic bearingassembly conduct the fluid to the thrust bearing and the circumferenceof the thrust plate 273. To this end, the dynamic pressure distributionhas a peak adjacent to the circumference of the thrust plate 273. Thelong and short dotted line 278 in FIG. 55 schematically shows thedynamic pressure distribution. The pump-out type hydrodynamic bearingassembly generates the peak dynamic pressure at the circumference so asto realize the robust rigidity against the disturbance motion when thefluid is supplied to the thrust member. The through-holes keeps thedynamic pressure above and under the thrust plate 273 even to stabilizethe rotation of the bearing assembly.

As illustrated in FIG. 55, in the structure of the complex radial-thrustbearing assembly wherein the radial bearing and the thrust bearing arecontinuously formed, the thrust bearing requires the fluid supplied fromthe radial bearing to the thrust bearing, to be enough for generatingthe dynamic pressure. The fluid is supplied from the upper end of theshaft through the radial gap between the shaft 271 and the sleeve 272during the rotation of the shaft 271. However, if the shaft 271 has acircular cross section as shown in FIG. 55, a sufficient amount of thefluid is hardly delivered to the trust plate 273. Also, the herringbonegrooves 274 in the radial bearing generates a peak of the radial dynamicpressure distribution at the middle portion of the radial bearing (atthe middle portion of the herringbones with the V-shaped indication inthe drawing). The fluid for generating this peak at the middle portionof the herringbones comes towards the radial bearing so that the fluidsupplied to the radial bearing is likely to be short even if thepump-out type spiral grooves 275 are provided on the thrust plate 273.When the fluid supplied to the thrust bearing is short, the dynamicpressure in the thrusting direction cannot be generated so that thesupporting force is also short. To this end, this causes the rotationalmember and the stationary member to be in contact with each other.

The need has been existed for a simple fastening mechanism forperpendicularly fastening the shaft with the thrust plate in a precisemanner, in order to improve the bearing rigidity. Some conventionalfastening methods and problems thereof will be described hereinafter.The fastening mechanism may bond the shaft directly with the thrustplate. This mechanism has a difficulty to keep the accuracy of theperpendicularity due to an uneven thickness of the adhesive. Thetolerance limits of the perpendicularity is 0.3 microns measured as thetilt relative to the thrust plate diameter of 20 millimeters. When theshaft has the diameter of 4 millimeters or more, it is hardly possibleto meet the tolerance limits even if the adhesive is cured while heldcorrectly.

As shown in FIG. 57(a), the fastening mechanism comprises a hollowcylindrical shaft 281, a thrust plate 283, and a volt extendingtherethrough for fastening the shaft and thrust plate. However, anuneven pressure biased by the bolt 286 or the washer 287 causes theelastic deformation of the shaft 281, thereby to result the malfunctionof the bearing assembly in this mechanism. Even a rubber pad attached tothe end of the hollow space of the shaft 281 provides the same result.The present inventors have found that the perpendicularity was 1.2micron due to the uneven pressure biased by the bolt 286 and it was 1.0micron with use of the rubber pad. It is hardly possible to reduce thedeviation of the radial component of the fastening force with use ofthis mechanism.

Also, another fastening mechanism comprises the cylindrical hollow shaft281, the thrust plate 283′, and a core member 288 secured on the thrustplate 283′ and connected with the shaft 281 by the shrink fitting.However, when cooled down to the room temperature, the outer surface ofthe shaft 281 radially expands with the elastic deformation so that theradial gap between the shaft and the sleeve has an unwanted influences.The present inventors have found that the outer diameter of the shaft281 radially expanded by 3 microns due to the shrink fitting and thebearing assembly could not practically be used.

Another prior arts fastening mechanism fastens the shaft and thrustplate by screwing a bolt onto the shaft 281 and the thrust plate 283.This fastening mechanism is simple and keeps the perpendicularity in aprecise manner. If the shaft 281 is made of stainless steel, then thisfastening mechanism can be used. However, if the shaft 281 is made ofceramics material, in which it is difficult to make a thread, themechanism can hardly be utilized in practical.

(Third Problem)

To address the third problem, i.e., the improvement of the activationfeature, many approaches have been proposed, for example by developingan effective spiral grooves in generating the thrust dynamic pressure.One of the solutions is reducing the floating rotation rate. Othersolutions include increasing the acceleration at the beginning of theactivation to minimize the time period in rotating with contact, andreducing the mass of the rotational member. Unless the rotation withcontact is avoided, at the beginning of the activation, a significantactivation torque is required and the bearing assembly wears quickly.Thus, the activation feature has to be further improved in future.

(Fourth Problem)

With respect to the fourth problem, i.e., an implementation of thecompact and lightweight bearing assembly, the hydrodynamic bearingassembly has achieved the improvement in comparison with the ballbearing assembly. However, in any event, the market still needs thebearing assembly to be more compact and lightweight, thus, a furtherimprovement is required.

Problems to be Solved

The conventional approaches for overcoming the half-whirl of the bearingassembly and for improving the bearing rigidity have each deficienciesas described above. Therefore, a purpose of the present invention is toprovide the hydrodynamic bearing assembly eliminating the half-whirl andimproving the rigidity against the disturbance.

Another purpose of the present invention is to provide the hydrodynamicbearing assembly improving the activation feature and satisfying thecompact and lightweight requirement.

SUMMARY OF THE INVENTION

The first aspect of the present invention is to provide the hydrodynamicbearing assembly eliminating the half-whirl and rotating at a highrotation rate in a stable manner. In particular, the hydrodynamicbearing assembly comprises a column shaft having an outer surfaceparallel to an axis; a hollow cylindrical sleeve having an inner surfacerotatably arranged around the outer surface of the shaft; and a radialbearing for generating a radial dynamic pressure due to a relativerotation between the sleeve and the shaft to keep them away from eachother; wherein a total radial gap of a pair of side radial gaps along adiameter defined between the outer surface of the shaft and the innersurface of the sleeve is approximately 3 microns or less, morepreferably approximately 2 microns or less. If the radial gap isdesigned to be extremely narrow, a fluid turbulence generated by thesurface roughness of the rotating and stationary member defining theradial bearing interrupts the continuous dynamic pressure distributionso as to prevent the half-whirl. Further, if the radial gap is extremelynarrow, in the complex radial-thrust bearing assembly, in which theradial and thrust bearings are combined, the relative tilt between thethrust plate and the thrust opposing surface in the thrust bearing canbe reduced to avoid the contact therebetween, thereby to improve thetilt rigidity.

Alternatively, a plurality of scratched notches are formed on thesurfaces defining the radial bearing so that a fluid turbulence isgenerated to avoid the half-whirl phenomenon. In particular, thehydrodynamic bearing assembly comprises: a column shaft having an outersurface parallel to an axis; a hollow cylindrical sleeve having an innersurface rotatably arranged around the outer surface of the shaft; and aradial bearing for generating a radial dynamic pressure due to arelative rotation between the sleeve and the shaft to keep them awayfrom each other; wherein either one of the outer surface of the shaftand the inner surface of the sleeve has a plurality of notches parallelto the axis, each notch extends from both ends or from positionsadjacent to the both ends of the bearing assembly, and has length of atleast one-fourth (¼) of that of the bearing assembly, wherein one to tenof the scratched notches are formed on the surface within an arc of 200microns, each notch has a depth within the range of approximately 1micron to approximately 20 microns and a width within the range ofapproximately 10 microns to approximately 200 microns, alternatively,wherein each scratched notch is formed on the surface with an intervalof an arc of at least 200 microns to another notch, each notch has adepth within the range of approximately 1 micron to approximately 20microns and a width within the range of approximately 200 microns toapproximately 500 microns.

The second aspect of the present invention is to provide thehydrodynamic bearing assembly, in which the bearing rigidity is improvedby forming a various configurations or grooves on the surfaces of themembers defining the radial bearing. Thus, the hydrodynamic bearingassembly comprises: a column shaft having an outer surface parallel toan axis; a hollow cylindrical sleeve having an inner surface rotatablyarranged around the outer surface of the shaft; and a radial bearing forgenerating a radial dynamic pressure due to a relative rotation betweenthe sleeve and the shaft to keep them away from each other; whereineither one of the outer surface of the shaft and the inner surface ofthe sleeve includes at lest one annular groove perpendicular to theaxis, the annular groove is such that a depth ratio defined by the depthof the annular groove relative to a diameter of the surface isapproximately 0.01 or less, and a width ratio defined by the width ofthe annular groove relative to a length of the bearing assembly isapproximately 0.2 or less. The peak of the radial dynamic pressuredistribution can be divided by providing the annular grooves so as toendure itself against the tilt motion. Also, the aforementioneddimension range of the annular grooves can minimize the reduction of thedynamic pressure.

Alternatively, the hydrodynamic bearing assembly comprises: a columnshaft having an outer surface parallel to an axis; a hollow cylindricalsleeve having an inner surface rotatably arranged around the outersurface of the shaft; and a radial bearing for generating a radialdynamic pressure due to a relative rotation between the sleeve and theshaft to keep them away from each other; wherein the outer surface ofthe shaft and the inner surface of the sleeve together defining a gaptherebetween are designed to have a configuration such that the gap iscontinuously and gradually enlarged substantially from a mid portion ofthe bearing assembly towards both ends along the axis when the shaft andthe sleeve are positioned concentrically, and such that the outersurface of the shaft is substantially parallel to the inner surface ofthe sleeve at the portions where the shaft and the sleeve are positionedmost closely because of a relative slope between the shaft and thesleeve. In order for the gap to continuously and gradually be enlargedsubstantially from the mid portion of the bearing assembly towards bothends along the axis, a hollow space of the sleeve may be continuouslyand gradually enlarged substantially from the middle portion of thebearing assembly towards both ends along the axis, or a diameter of theshaft may be continuously and gradually shrunk substantially from themiddle portion of the bearing assembly towards both ends along the axis.Also, both approaches may be combined.

An another approach to keep the thrust gap in the thrust bearing to beconstant is to keep the relative dimensions between the radial gap andthe thrust gap to fall within the particular range. Therefore, thehydrodynamic bearing assembly comprises: a radial bearing including acolumn shaft having an outer surface parallel to an axis, and a hollowcylindrical sleeve having an inner surface rotatably arranged around theouter surface of the shaft, the radial bearing for generating a radialdynamic pressure due to a relative rotation between the sleeve and theshaft; and a pair of thrust bearings, each including at least one thrustplate formed or secured onto either one of the shaft and the sleeve, anda pair of thrust opposing surfaces formed or secured onto the other oneof the shaft and the sleeve, the thrust bearing for generating a thrustdynamic pressure due to the relative rotation between the thrust plateand the thrust opposing surface; wherein a total radial gap F of a pairof side radial gaps along a diameter in the radial bearing, a length Lof the bearing assembly along the axis, and an outer diameter G in thethrust bearing satisfy the following condition; (F/L)<(D/G).Alternatively, the total radial gap F, and a total thrust gap D satisfythe following condition; kF<D, wherein k is constant in the range of 2to 10.

The third aspect of the present invention is to provide the hydrodynamicbearing assembly, in which the bearing rigidity is improved by providinga pump-out type hydrodynamic bearing assembly and by supplying asufficient amount of the fluid with the pump-out type hydrodynamicbearing assembly. Thus, the hydrodynamic bearing assembly comprises: aradial bearing including a column shaft having an outer surface parallelto an axis, and a hollow cylindrical sleeve having an inner surfacerotatably arranged around the outer surface of the shaft, the radialbearing for generating a radial dynamic pressure due to a relativerotation between the sleeve and the shaft; and a thrust bearingincluding a thrust plate formed or secured onto either one of the shaftand the sleeve, and a thrust opposing surface formed or secured onto theother one of the shaft and the sleeve, the thrust bearing for generatinga thrust dynamic pressure due to the relative rotation between thethrust plate and the thrust opposing surface; wherein the thrust bearingis a pump-out type, and conducts the fluid in the thrust bearing in thedirection from the axis to circumference, and wherein at least onelongitudinal groove formed on either one of the outer surface of theshaft and the inner surface of the sleeve, the longitudinal grooveextending parallel or offset to the axis through both ends along theaxis.

Alternatively, the pair of the thrust bearings connect with the radialbearing, the thrust gap is kept constant, and at least one of thrustbearings is the pump-out type one, so that a more robust rigidity can berealized. In particular, the hydrodynamic bearing assembly comprises: acolumn shaft having a pair of small columns concentrically formed on endsurfaces; a hollow cylindrical sleeve rotatably arranged around an outersurface of the shaft parallel to the axis of the shaft; and a pair ofdonut-shaped thrust plates arranged on both ends of the sleeve, each ofthe thrust plate including a through-hole, through which the smallcolumn extends; a radial bearing defined between the outer surface ofthe shaft and the inner surface of the sleeve; and a thrust bearingdefined between the pair of thrust plates and thrust opposing surfacesopposing to the thrust plates; wherein each thrust bearing communicatingin fluid with the radial bearing, and grooves are formed on either oneof the thrust plate and thrust opposing surface to generate a thrustdynamic pressure due to a relative rotation between the thrust plate andthrust opposing surface, and wherein the grooves of at least one of thethrust bearings being a pump-out type groove for conducting the fluid inthe direction from the axis to a circumference thereof.

The fourth aspect of the present invention is to provide thehydrodynamic bearing assembly, in which the bearing rigidity is improvedby increasing the perpendicularity between the shaft and the thrustplate and by fastening thereof. In particular, the hydrodynamic bearingassembly comprises: a radial bearing including a hollow cylindricalshaft having an outer surface parallel to an axis, a hollow cylindricalsleeve having an inner surface rotatably arranged around the outersurface of the shaft, the radial bearing for generating a radial dynamicpressure due to a relative rotation between the sleeve and the shaft;and a thrust bearing including a disk-shaped thrust plate securedperpendicularly onto one end of the axis of the shaft the thrust platehaving a through-hole formed concentrically to the axis, and a thrustopposing surface of one end surface of the sleeve, opposing to thethrust plate, the thrust bearing for generating a thrust dynamicpressure due to the relative rotation between the thrust plate and thethrust opposing surface; a constraint member closely fit within theinner surface of a hollow space of the shaft for securing the shaftthereto; and a fastening member having means for engaging with theconstraint member; wherein the engaging means extends through thethrough-hole of the thrust plate to engage with the constraint member sothat the shaft and the sleeve are secured. The constraint member mayhave a cylindrical extension member extending along the axis beyond anend surface of the thrust plate for closely fitting within thethrough-hole of the thrust plate. The engaging means may include a malescrew extending from either one of the constraint member and thefastening member and a female screw.

The fifth aspect of the present invention is to provide the hydrodynamicbearing assembly, in which the endurance and the reliablity thereof areimproved by improving the activation feature. In particular, thehydrodynamic bearing assembly comprises: a radial bearing including acolumn shaft having an outer surface parallel to an axis, and a hollowcylindrical sleeve having an inner surface rotatably arranged around theouter surface of the shaft, the radial bearing for generating a radialdynamic pressure due to a relative rotation between the sleeve and theshaft; and a thrust bearing including a thrust plate formed or securedonto either one of the shaft and the sleeve, and a thrust opposingsurface formed or secured onto the other one of the shaft and thesleeve, the thrust bearing for generating a thrust dynamic pressure dueto the relative rotation between the thrust plate and the thrustopposing surface; wherein the radial bearing connects with the thrustbearing adjacent connecting portions forming conduits for a fluid, thefluid is communicated via a the conduit having the corn portions or thecontinuous and smooth curve portions without bending points, a first andsecond distances m and n defined between two symmetry ascendingpositions to the axis of the corn or curve portion of the sleeve and thethrust plate, respectively, satisfy a following condition; m<n, so thatthe radial bearing and the thrust bearings adjacent connecting portionsare formed as corn portions or continuous and smooth curve portionswithout bending points.

An another approach to keep the thrust gap in the thrust bearing to beconstant is to provide the hydrodynamic bearing assembly comprises: aradial bearing including a column shaft having an outer surface parallelto an axis, and a hollow cylindrical sleeve having an inner surfacerotatably arranged around the outer surface of the shaft, the radialbearing for generating a radial dynamic pressure due to a relativerotation between the sleeve and the shaft; and a pair of thrustbearings, each including at least one thrust plate formed or securedonto either one of the shaft and the sleeve, and a pair of thrustopposing surfaces formed or secured onto the other one of the shaft andthe sleeve, the thrust bearing for generating a thrust dynamic pressuredue to the relative rotation between the thrust plate and the thrustopposing surface; wherein the radial bearing connects with the pair ofthe thrust bearings adjacent connecting portions forming conduits for afluid, the radial bearing and the thrust bearings adjacent connectingportions are formed as corn portions or continuous and smooth curveportions without bending points, and wherein, when the shaft and thesleeve are positioned so that they are concentric and each thrust gap isthe same, a one-side radial gap F/2 along the radius in the radialbearing, a one-side thrust gap D/2 in the thrust bearing, and a gapdistance S/2 of the conduits adjacent connecting portions satisfy thefollowing condition; F<S<D.

The sixth aspect of the present invention is to provide the compact andlightweight hydrodynamic bearing assembly. In particular, thehydrodynamic bearing assembly includes opposing portions among theshaft, the sleeve, the thrust plate, and at least one of the thrustopposing surfaces, which are made of ceramics material. The ceramicsmaterial is selected from a group consisting of alumina, zirconia,silicon carbide, silicon nitride, and sialon.

The seventh aspect of the present invention is to provide a spindlemotor as well as a memory device and bar code reader incorporating thespindle motor, which are able to rotate in a stable manner and arerobust against the external motions.

BRIEF DESCRIPTION OF DRAWINGS

FIG. 1 is a perspective view of one embodiment of the hydrodynamicbearing assembly according to the present invention.

FIG. 2 is a perspective view of the alternative sleeve of thehydrodynamic bearing assembly in FIG. 1.

FIG. 3 is a perspective view of another embodiment of the hydrodynamicbearing assembly according to the present invention.

FIG. 4 is a side view of the conventional hydrodynamic bearing assembly.

FIG. 5 is a side view of the another embodiment of the hydrodynamicbearing assembly according to the present invention.

FIG. 6 is a schematic view of an evaluation equipment for a translationrigidity and a tilt rigidity of the hydrodynamic bearing assembly.

FIG. 7 is a graph showing an evaluation index with use of the evaluationequipment of FIG. 6.

FIGS. 8(a) to 8(d) are cross sectional views of the shaft illustratingthe direction of the external force relative to the longitudinalgrooves.

FIG. 9 is a cross sectional view of another embodiment of thehydrodynamic bearing assembly according to the present invention.

FIG. 10 is a cross sectional view of the hydrodynamic bearing assemblyin FIG. 9 during the rotation thereof.

FIG. 11 is a cross sectional view of the conventional hydrodynamicbearing assembly during the rotation thereof.

FIG. 12 is a cross sectional view of the hydrodynamic bearing assemblyin FIG. 9 illustrating various dimensions thereof.

FIG. 13 is a cross sectional view of further another embodiment of thehydrodynamic bearing assembly according to the present invention.

FIG. 14 is a cross sectional view of the hydrodynamic bearing assemblyin FIG. 13 during the rotation thereof.

FIGS. 15(a) to 15(c) are cross sectional views of the anotherembodiments of the hydrodynamic bearing assembly according of thepresent invention, illustrating a couple of combinations of the shaftand sleeve.

FIG. 16 is a cross sectional view of the spindle motor incorporating thehydrodynamic bearing assembly in FIG. 9.

FIG. 17 is a cross sectional view of the spindle motor incorporating thehydrodynamic bearing assembly according to further another embodiment ofthe present invention.

FIG. 18 is a cross sectional view of the hydrodynamic bearing assemblyin FIG. 17, illustrating the dynamic pressures in the bearings.

FIG. 19 is a cross sectional view of the alternative hydrodynamicbearing assembly in FIG. 17.

FIG. 20 is a cross sectional view of another embodiment of thehydrodynamic bearing assembly according to the present invention.

FIG. 21 is a side view of the alternative hydrodynamic bearing assemblyin FIG. 20.

FIG. 22 is side view of the further alternative hydrodynamic bearingassembly in FIG. 20.

FIG. 23 is a cross sectional view of the spindle motor incorporating thehydrodynamic bearing assembly according to further another embodiment ofthe present invention.

FIG. 24 is a cross sectional view of the hydrodynamic bearing assemblyin FIG. 23 during the rotation thereof.

FIGS. 25(a) and 25(b) are cross sectional views of the alternativehydrodynamic bearing assembly in FIG. 23 during the rotation thereof.

FIGS. 26(a) to 26(d) are cross sectional views of the furtheralternative hydrodynamic embodiment of the bearing assembly of FIG. 23during the rotation thereof.

FIG. 27 is a cross sectional view of even further alternative embodimentof the hydrodynamic bearing assembly in FIG. 23 during the rotationthereof.

FIG. 28 is a cross sectional view of the spindle motor incorporating thehydrodynamic bearing assembly in FIG. 27.

FIG. 29 is a cross sectional view of even further alternative embodimentof the hydrodynamic bearing assembly according to the present invention.

FIGS. 30(a) and 30(b) are perspective views of the constraint memberconstituting the hydrodynamic bearing assembly in FIG. 29

FIG. 31 is a cross sectional view of the alternative embodiment of thehydrodynamic bearing assembly in FIG. 29.

FIG. 32 is a cross sectional view of the further alternative embodimentof the hydrodynamic bearing assembly according to the present invention.

FIG. 33 is a cross sectional view of the hydrodynamic bearing assemblyin FIG. 32.

FIG. 34 is a cross sectional view of the further alternative embodimentof the hydrodynamic bearing assembly according to the present invention.

FIG. 35 is a cross sectional view of the further alternative embodimentof the hydrodynamic bearing assembly according to the present invention.

FIG. 36 is an enlarged view of the hydrodynamic bearing assembly in FIG.35 when halted.

FIG. 37 is a cross sectional view of the alternative embodiment of thehydrodynamic bearing assembly in FIG. 35.

FIG. 38 is a cross sectional view of the alternative embodiment of thehydrodynamic bearing assembly in FIG. 35.

FIG. 39 is a cross sectional view of the alternative embodiment of thehydrodynamic bearing assembly in FIG. 35.

FIGS. 40(a) to 40(c) are cross sectional views of the spindle motoraccording to the present invention, illustrating the rotor magnet andthe electromagnet.

FIG. 41 is a cross sectional view of the alternative embodiment of thehydrodynamic bearing assembly according to the present invention.

FIG. 42 is a cross sectional view of the conventional spindle motor.

FIG. 43 is a perspective view of the thrust plate used for thehydrodynamic bearing assembly in FIG. 42.

FIG. 44 is a transverse cross sectional view of the shaft, illustratingthe half-whirl of the hydrodynamic bearing assembly.

FIG. 45 is a cross sectional view of the hydrodynamic bearing assembly,illustrating a dynamic pressure distribution of the hydrodynamic bearingassembly.

FIG. 46 is a cross sectional view of the hydrodynamic bearing assembly,illustrating another dynamic pressure distribution of the hydrodynamicbearing assembly.

FIG. 47 is a cross sectional view of the hydrodynamic bearing assembly,illustrating further another dynamic pressure distribution of thehydrodynamic bearing assembly.

FIG. 48 is a vertical cross sectional view of another conventionalhydrodynamic bearing assembly.

FIG. 49 is a transverse cross sectional view of the hydrodynamic bearingassembly in FIG. 48.

FIG. 50 is a partially fragmentary perspective view of anotherconventional hydrodynamic bearing assembly.

FIG. 51 is an enlarged view of the sleeve of the hydrodynamic bearingassembly in FIG. 50.

FIG. 52 is a side cross sectional view of another hydrodynamic bearingassembly used for the conventional bar code reader.

FIG. 53 is a cross sectional view of the hydrodynamic bearing assemblyin FIG. 52, illustrating the dynamic pressure distribution thereof.

FIG. 54 is a transverse cross sectional view of a conventional biasingmeans.

FIG. 55 is a side cross sectional view of another embodiment of theconventional hydrodynamic bearing assembly.

FIG. 56 is a partial top view of the hydrodynamic bearing assembly inFIG. 55.

FIGS. 57(a) and 57(b) are side cross sectional views of further anotherembodiment of the conventional hydrodynamic bearing.

PREFERRED EMBODIMENTS OF THE INVENTION

The first aspect of the present invention is to address overcoming ahalf-whirl. As discussed above with reference to FIG. 44, the half-whirlis a phenomenon where the rotational member (either one of the shaft andsleeve) whirls relatively to the stationary member (the other one of theshaft and sleeve). In order to achieve the relative rotation withoutcontact between the rotational member and the stationary member, acertain gap is essential between the opposing surfaces of both members.It is understood that the half-whirl phenomenon is theoreticallyunavoidable under the existence of the gap,

First Embodiment

The first embodiment of the hydrodynamic bearing assembly according tothe present invention, which mainly address overcoming the half-whirl,will be described hereinafter. The present inventors have found nohalf-whirl phenomenon appeared if a radial gap between the outer surfaceof a shaft and the inner surface of a sleeve, both of which define aradial bearing, is extremely narrow. For example, the difference betweenthe outer diameter of the shaft and the inner diameter is preferablyabout 3 microns or less and more preferably about 2 microns or less sothat no half-whirl phenomenon is observed. There seems some reasons asfollows. The outer surface of a shaft or the inner surface of a sleevestill has a certain roughness remaining even after the smootheningprocess. This surface roughness causes a turbulence of the fluid betweenthe rotating and stationary member. As discussed above, the continuousdynamic pressure distribution causes the half-whirl. It is understoodthat the fluid turbulence, which is greater as the radial gap isnarrower, interrupts the continuous dynamic pressure distribution sothat the half-whirl is prevented. The embodiment relates to thehydrodynamic bearing assembly, in which the total radial i.e., thedifference between the outer diameter of the shaft and the innerdiameter, is preferably about 3 microns or less and more preferablyabout 2 microns or less. Although the radial gap may be designed to bezero theoretically, then there is a high possibility that the sleevecannot be arranged around the shaft or no rotation can be possible dueto the manufacturing accuracy of the sleeve and the shaft. To avoid suchpossibility, the gap is designed to be more than about 1 micron.

When the radial gap is extremely narrow as described above, it isunderstood that the outer surface of the shaft and the inner surface ofthe sleeve keep in contact with each other while the spindle motor ishalted. If both of the opposing surfaces have such roughness, then thefriction between the opposing surfaces is greater. This causes the shaftand the sleeve to wear quickly, and the greater torque is required tostart rotating the spindle motor. The increased torque due to thefriction of the opposing contact surfaces can be avoided by furthersmoothening either one of the opposing surfaces so that the frictionbetween the opposing surfaces is reduced. Also, even when one of theopposing surfaces is smoothened considerably, the other one of theopposing surfaces may have the appropriate roughness remaining togenerate the fluid turbulence in the dynamic pressure distribution forpreventing the half-whirl.

When the outer surface of the shaft is further smoothened, it has thesurface roughness Rp, preferably about 0.2 micron or less, or morepreferably about 0.15 micron or less. On the other hand, when the innersurface of the sleeve is further smoothened, taking account of thedifficulty to smoothen, it has the surface roughness Rp, preferablyabout 0.4 micron or less, or more preferably about 0.35 micron or less.The surface roughness Rp is representative of the deviation from theaverage of the roughness curve to the measured curve. Although thesurface roughness Rp may be designed to be zero theoretically, it isdetermined based upon the restriction condition of the smootheningaccuracy and the cost for smoothening. In addition, in order to ensurethe narrow radial gap between the shaft and the sleeve according to thepresent invention, preferably, the circularity deviation of the crosssections of the shaft and the sleeve are taken into consideration. Theouter surface of the shaft has the circularity deviation preferably ofabout 0.2 microns or less, or more preferably of about 0.12 microns orless. Similarly, the inner surface of the sleeve has the circularitydeviation preferably of about 0.7 microns or less, or more preferably ofabout 0.4 microns or less, taking account of the difficulty to produce.Although the circularity deviation may be designed to be zerotheoretically, it is also determined based upon the restrictioncondition of the production and the manufacturing cost.

Second Embodiment

The second embodiment of the hydrodynamic bearing assembly according tothe present invention, which addresses overcoming the half-whirl, willbe described hereinafter. The hydrodynamic bearing assembly of theembodiment includes a plurality of scratched notches as means forovercoming the half-whirl, instead of the conventional notches formedeither on the outer surface of the shaft or the inner surface of thesleeve. The scratched notches generate the fluid turbulence interruptingthe continuous dynamic pressure distribution for eliminating thehalf-whirl.

FIGS. 1 and 2 illustrate the hydrodynamic bearing assembly of theembodiment. The hydrodynamic bearing assembly comprises a shaft 1, asleeve 2 arranged around the shaft 1 as indicated by the phantom line,and a thrust plate 3 perpendicularly secured to the shaft 1 opposing tothe bottom surface (referred to as a thrust opposing surface) of thesleeve 2. The thrust plate 3 includes a plurality of spiral groovesopposing to the thrust opposing surface of the sleeve 2 for generatingthe thrust dynamic pressure. When the sleeve 2 rotates relative to theshaft 1 and the thrust plate 3, the radial dynamic pressure and thethrust dynamic pressure are generated between the sleeve 2 and the shaft1, and between the thrust plate 3 and the thrust opposing surface of thesleeve 2, respectively. Such generated dynamic pressures cause thesleeve 2 to rotate relative to the shaft 1 and the thrust plate 3without any contact therebetween.

As shown in FIG. 1, the outer surface of the shaft 1 include thescratched notches 6, 7. In particular, each of the scratched notches 6,extends across and along the shaft 1, while each of the scratchednotches 7 extends along the shaft 1 but is interrupted in part adjacentthe mid portion of the sleeve 2. Each of the scratched notches 6 and 7is much finer than the conventional notch 221 indicated in FIG. 48,which for example, has a depth in the range of about 1 micron to about20 microns, and a width in the range of about 10 microns to about 500microns. No problem was found in operation for the scratched notcheswith the depth of about 20 microns or less, while the depth may beinfluenced by the scratching process.

The finer scratched notches 6 and 7 formed on the outer surface of theshaft 1 cause the turbulence of the fluid in the radial bearing so as toeliminate the half-whirl. Also, the scratched notches 6 and 7 are sofine that the radial dynamic pressure is maintained in the radialbearing. Therefore, the hydrodynamic bearing assembly of the embodimentraises no problem that the prior art suffers, for example, the reductionof the translation rigidity against the disturbance motion along apredetermined direction, and the increase of the risk of the rotationwith contact.

Contrary, each of the scratched notches 7 formed along but not acrossthe shaft 1 extends preferably from the bottom end and/or the top end,or portions adjacent thereto, upwardly and/or downwardly to the middleportion of the shaft 2, respectively. Also, each of the scratchednotches 7 has a longitudinal length which is more than one-fourth of theentire length (thus in total, half of the entire length) of the shaft 2.Since the scratched notches 6 and 7 extends from the bottom end and/orthe top end of the shaft 2 to cause the inside of the notches to be opento the atmosphere, the continuous dynamic pressure distribution causingthe half-whirl can advantageously be interrupted. The prior art(Japanese Patent Laid-Open Publication 2-150504) fails to show thisadvantage, it discloses a plurality of longitudinal bands in thedirection of the bearing axis, which includes circumferential microground streaks formed on the inner surface of the sleeve 202.

To eliminate the half-whirl satisfactorily, preferably, about one to tenof the scratched notches 6 and 7 having the width in the range of about10 microns to about 200 microns is formed within every 200 microns ofthe length in the direction of the circumference of the shaft 1.Similarly, one of the scratched notches 6 and 7 having the width in therange of about 200 microns to about 500 microns is preferably formedwith at intervals of at most 200 microns in the direction of thecircumference of the shaft 1. The expression of “with at most 200microns of an interval to another” means that a plurality of scratchednotches with the width in the range of about 200 microns to about 500microns are arranged away from each other with a interval of at most 200microns or, preferably less than it. The experiments conducted by thepresent inventors revealed that when the scratched notches include thedepth in the range of one to three microns are arranged away from eachother with a space of 200 microns (five notches per one millimeter), thebearing rotation can be implemented in a stable manner.

The scratched notches do not have to be formed on the outer surface ofthe shaft at a constant interval. Rather, the scratched notches arenecessarily formed on the longitudinal width of about more thanone-tenth of the circumference in order to prevent the half-whirl. Forexample, when the circumference of the shaft is 10 millimeters, thelongitudinal width where the scratched notches are formed is about morethan one millimeter. Also, the scratched notches may be divided on threeseparate circumferencial portions at the regular interval. While each ofthe scratched notches 6 and 7 are illustrated as a single continuousfine notch in FIG. 1, such a scratched notch is formed, for example, bypressing the shaft on the rubstone and moving thereof in the axisdirection. Alternatively, a group of the intermittent scratched notchesalong the axis direction may be formed, for example by using a annularrubstone, rotating the shaft around its own axis, and movingtherethrough. The axis direction may not be perpendicular to the axis ina precise manner and may be inclined to some extent. This applies forthe case indicated in FIG. 2.

FIG. 2 illustrates the sleeve 2 including the inner surface on which thescratched notches 6 and 7 are formed. The scratched notches 6 and 7 ofFIG. 2 are formed in a similar manner as those of FIG. 1. While theconventional streak band comprises circumferential micro ground streaks,the scratched notches 6 and 7 of FIG. 2 are formed longitudinally alongthe sleeve 2. Therefore, advantageously, the scratched notches 6 and 7are readily formed, for example by a couple of longitudinal movement ofa cylindrical rubstone with an appropriate diameter through the sleeve 2without rotation. In addition, even the sleeve has a small diameter (forexample, 2 millimeters), still it can readily be processed.

FIGS. 1 and 2 illustrate that the shaft 1 is the stationary member andthe sleeve 2 is the rotational member, however, the present invention isnot limited thereto, thus, the shaft 1 may be the rotational member andthe sleeve 2 may be the stationary member. In addition, in the drawings,the combination of the scratched notches 6 and 6 a extending along andacross the shaft 1 and the scratched notches 7 and 7 a extending alongbut partially being interrupted in the mid portion of the shaft 1 areillustrated. However, this is so illustrated just for clarification, andany combinations are acceptable without departing from the scope of thepresent invention.

The half-whirl is prevented by providing the scratched notches accordingto the embodiment on either one of the outer surface of the shaft andthe inner surface of the sleeve, even when both of the opposing surfacesare further smoothened. To this end, according to the embodiment, thefriction between both of the opposing surfaces can be further decreased,and the activation torque can also be reduced.

The second aspect of the present invention is to address improving thebearing rigidity.

Third Embodiment

The third embodiment of the hydrodynamic bearing assembly according tothe present invention, which addresses improving the bearing rigidity,will be described hereinafter. The third embodiment relates to therestriction conditions in the configurations and the dimensions of thelongitudinal grooves extending along the axis direction as well as theannular grooves perpendicularly crossing the longitudinal grooves. Asdiscussed above, it is well known in the art that in order to improvethe tilt rigidity as shown in FIG. 3, the annular groove 8 is providedon the outer surface of the shaft 1 at the middle portion in the axisdirection, extending substantially perpendicular to the axis directionacross the longitudinal grooves for supporting the sleeve 2 at variouspoints. Meantime, the annular grooves 8 reduce the translation rigiditycausing further disadvantages as described above. According to theembodiment, the translation rigidity can advantageously be improved,while maintaining the tilt rigidity, by optimizing the width b and thedepth t of the annular groove 8. The annular groove 8 has a depth ratio(which is referred to as a ratio of the depth of the annular groove 8relative to the diameter of the shaft 1) of 0.01 or less, and a widthratio (which is referred to as a ratio of the width of the annulargroove 8 relative to the longitudinal length of the sleeve 2) of 0.2 orless. If the depth ratio is 0.01 or more, then the total volume of thegap in the bearing assembly is excessive to cause the reducedtranslation rigidity. If the width ratio is 0.2 or more, then thetilt-rigidity is reduced. If the depth ratio is 0.0001 or less, or ifthe width ratio is 0.01 or less, then the peak of the dynamic pressuredistribution is not divided due to the annular groove 8.

As discussed above, in the hydrodynamic bearing assembly having theshaft 1 with the outer surface on which longitudinal groove 9 is formedalong the axis as shown in FIG. 4, if the disturbance motion is appliedin the direction aligning to the longitudinal grooves 9, then thetranslation rigidity is reduced. In order to overcome the problem, theupper grooves 9 a are shifted along the rotating direction relative tothe lower groove 9 b so that the dependency of the translation rigiditycan be reduced.

The reasons leading this conclusion will be described hereinafter.Several samples were prepared varying the dimension of the annulargrooves 8 and position of the longitudinal grooves 9, and then wereevaluated for the variation with respect to the translation rigidity andthe tilt rigidity due to the external force.

FIG. 6 is a schematic view of an evaluation equipment. The evaluationequipment comprises a rotor hub 11 incorporated with the hydrodynamicbearing assembly secured on the base plate 10, an arc nozzle 12 having awing at the tip for guiding air adjacent to and around rotor hub 11, anda capacitance prove 13 located on the opposite side over the rotor hub11. The base plate 10 has a supporting column 14 at the end forsupporting the arc nozzle 12. The elevation H1 of the arc nozzle 12 canbe adjusted by a screw 15. The arc nozzle 12 includes a guiding hole 16to which compressed air is guided from a compressor through ahigh-pressure hose (not shown). The external force P to be applied withthe rotor hub 11 can be varied by adjusting the air pressure. Meantime,the capacitance prove 13 is secured with an L-shaped flange 17, which isarranged on the base plate 10 at the other end thereof. The elevation H2of the capacitance prove 13 can also be adjusted by sliding the flange17 within the longitudinal hole 18 and by fastening it with a nut 19.The capacitance prove 13 measures the variation of distance (Δx) to theouter surface of the rotor hub 11. In measuring the translationrigidity, the arc nozzle 12 is positioned on one side of the rotor hub11 and the capacitance prove 13 is located on the other side.

FIG. 7 is a graph having a vertical axis representative of the externalforce P and horizontal axis representative of the variation of distance(Δx). If the external force P is applied with the sample (i) and (ii),and the variations of distance (Δx) of the samples (i) and (ii),respectively, as shown in FIG. 7, then the rigidity of the sample (i) isdetermined greater than that of sample (ii). The gradients of the lines(i) and (ii) are indicated as an indexes in Tables 2 and 3.

The shaft 1 and the sleeve 2, which have standard dimensions withcircularity deviation and cylindricality finished with the accuracy ofthe class 4 defined by the JIS (Japanese Industrial Standard) B 0401“System of Limits and Fits”, were prepared. Then, the shaft 1 and thesleeve 2 were selected and fit together to make a hydrodynamic bearingassembly with the total radial gap along the diameter of 5 microns orless. The annular groove 8 was provided substantially in the mid portionof the shaft 1. The specification of the annular groove 8 is indicatedas dimension ratios of Sample Nos. 1-1 to 1-10. Also in Table 1, thetranslation rigidity and the tilt rigidity influenced by the annulargroove 8 alone without forming the longitudinal groove 9 are indicatedas the gradients of FIG. 7, which were obtained by means of theevaluation equipment of FIG. 6. The alphabet L and J are representativeof the longitudinal length along the axis direction of the bearingassembly and the diameter of the shaft 1, respectively. TABLE 1 Widthratio of Depth ratio of Sample annular groove annular groove Trans. TiltNo. (b/L) (t/J) rigidity rigidity *1-1 No annl. grv. No annl. grv. 1 0*1-2 0.3 0.1 0.05 0.2 *1-3 0.3 0.01 0.08 2 *1-4 0.3 0.005 0.1 3 *1-5 0.20.1 0.06 2  1-6 0.2 0.01 0.8 10  1-7 0.2 0.005 0.9 15 *1-8 0.1 0.1 0.082  1-9 0.1 0.01 0.8 8  1-10 0.1 0.005 1.0 20*comparative data

As can be seen from Table 1, if the width ratio of the annular groove 8is 0.2 or more, the translation rigidity is noticeably reduced becauseof the increase the total volume of the gap in the bearing assembly.However, if the depth ratio of the annular groove 8 is 0.01 or less,then the translation rigidity is substantially improved. Apparently, thetilt rigidity is improved due to the existence of the annular groove 8but insufficiently. As can be seen from Table 1, if the width ratio ofthe annular groove 8 is set to be 0.01 or less, and if the depth ratioof the annular groove 8 is set to be 0.2 or less, then the tilt rigiditycan be remarkably improved without reducing the translation rigidity bythe synergism effect.

Next, the present inventors investigated how the position of the numbersof the longitudinal grooves 9, and the direction of the external forcegives impact on the translation rigidity and the tilt rigidity. Foursamples of the shaft 2 having various longitudinal grooves 9 wereprepared. FIG. 8 illustrates the cross section of those samples of theshafts. Prepared were the first Sample 2-1, FIG. 8(a) having only onelongitudinal groove 9 extending across the annular groove 8 as shown inFIG. 4, the second Sample 2-2, FIG. 8(b) having three longitudinalgroove 9 extending across the annular groove 8 with a regular angularinterval, the third Sample 2-3, FIG. 8(c) having an upper longitudinalgroove 9 a and a lower longitudinal groove 9 b shifted by 90 degreesrelative to the upper one as shown in FIG. 5, and the fourth Sample 2-4,FIG. 8(d) having three upper longitudinal grooves 9 a and three lowerlongitudinal grooves 9 b with a regular angular interval and shifted by60 degrees relative to the upper ones. Each sample has the identicalannular groove 8 provided substantially in the mid portion having thedepth ratio (the depth relative to the diameter of the shaft) of 0.01 orless and the width ratio (the width relative to the diameter of theshaft) of 0.2 or less. Also, each sample has the longitudinal groovehaving the same rectangular cors section and the same width ratio (thewidth relative to the diameter of the shaft) of 0.005. With use of theevaluation equipment of FIG. 6, the translation rigidity and the tiltrigidity were measured with the index (gradient) shown in FIG. 7, whenthe external force was applied to the directions as indicated by arrowsin FIG. 8. TABLE 2 Trans. Rigidity Tilt Rigidity Sample No. A B A B *2-1(a) 1.0 0.6 16 15 *2-2 (b) 1.0 0.5 17 16  2-3 (c) 0.9 0.9 20 18  2-4 (d)0.8 0.8 18 17*comparative data

As can be seen, the comparative samples (2-1) and (2-2) have thetranslation rigidity against the external force in the A directiongreater than that in the B direction, although the samples (2-3) and(2-4) according to the embodiment have the same translation rigidityagainst the external force in the A and B directions. Also, the samples(2-3) and (2-4) advantageously have the tilt rigidity against theexternal forces in both directions greater than those of the comparativesamples (2-1) and (2-2) because the position of the gaps are radiallyspread.

The longitudinal grooves 9 can be formed with use of any appropriateprocesses such as the plasma etching, the shotblasting, thelaser-abrading, the grinding, and the turn-chiseling. Although theembodiment has been discussed as the annular grooves 8 and thelongitudinal grooves 9 are provided on the outer surface of thestationary member of the shaft 1, it should be understood that thesimilar effect can be expected if the annular grooves 8 and thelongitudinal grooves 9 are provided on the inner surface of the sleeve 8when the sleeve 2 is the stationary member.

Fourth Embodiment

The fourth embodiment of the hydrodynamic bearing assembly according tothe present invention, which mainly address improving the bearingrigidity, will be described hereinafter, with reference to drawings.FIG. 9 is a cross section of the bearing assembly of the embodiment,illustrating the shaft 1 and the sleeve 2 a rotatably arranged aroundthe shaft 1. According to the embodiment, while the shaft 1 is straightand has the transverse circular cross section as being even along thebearing axis, the sleeve 2 a has the transverse cross section with aninner diameter continuously increasing towards to the upper and lowerends.

In FIG. 8, when the bearing sleeve 2 a has the length L along the axis,the gap between the shaft 1 and the sleeve 2 a is the minimum at theposition of L/2 from the both edges of the sleeve 2 a. When the leftside gap and the right side gap between the shaft 1 and the sleeve 2 ahave the sizes of f1 and f2 at the position of L/2 from the edges, thetotal gap F at the same position is;F=f1+f2If the shaft 1 and sleeve 2 a are arranged concentrically to each other,then the f1 equals to f2 (f1=f2), which is referred to as a one-sideminimum gap f. Thus, the total gap F is 2f (F=2f). The embodimentadvantageously has the total gap F to be minimum at the position of L/2from the edges, the position causing the total gap F to be minimum isunnecessarily the exact mid portion, rather it may be adjacent to themid portion.

The gaps at the end positions of the sleeve 2 a is the maximum, and theleft side maximum gap and the right side maximum gap between the shaft 1and the sleeve 2 a have the sizes of f1max and f2max at the endpositions.f1max=f1+αwherein α stands for the increased gap relative to that at the midposition. If the shaft 1 and sleeve 2 a are arranged concentrically toeach other, then the one-side maximum gap fmax at the end positions is;fmax=f+α

FIG. 10 shows the shaft 1 and sleeve 2 a of FIG. 9 during the rotation.The shaft 1 is inclined relative to the shaft 2 a because of an externalbiasing force as will be described below. The gaps become narroweradjacent left upper portion M1 and the right lower portion M2 in FIG.10, when the spindle motor includes shaft 1 rotating inside the sleeve 2a. The narrower gap generates the higher dynamic pressure, thereby tokeep the shaft 1 and the sleeve 2 a away from each other during therotation. This is also true for the spindle motor including sleeve 2 arotating around the shaft 1.

The parallel lines indicated adjacent to the end positions M1 and M2 inFIG. 10, schematically illustrate the dynamic pressure distribution. Inorder to minimize the whirl of the shaft 1 of the motor during therotation, the dynamic pressures generated adjacent to the end positionsM1 and M2 have to be increased across the extensive area by designingthe configuration of the outer surface of the sleeve 2 a so that it hastangential lines substantially parallel to the outer surface of theinclined shaft 1 across the wide area.

As can be more clearly understood in comparison with the conventionalbearing assembly of FIG. 11, the dynamic pressure adjacent to the endpositions M1′ and M2′ are poorly or unsatisfactorily generated as shownby the parallel lines in FIG. 11, when the shaft 1 is inclined relativeto the sleeve 2 and the edge portions of the shaft 1 advance towards thesleeve 2. Because of the insufficient dynamic pressure, when the shaft 1is externally biased to tilt relative to the sleeve 2 during therotation, the sleeve 2 is likely contact with the shaft 1 andeventually, a good endurance can hardly be expected for the bearingassembly.

The configuration of the inner surface of the sleeve 2 a of the bearingassembly according to the embodiment, which the present inventors havefound advantageous, will be described hereinafter, with reference to thedrawings. FIG. 12 is a drawing similar to FIG. 9, but is turned by 90degrees. The shaft 1 and the sleeve 2 a are arranged concentrically. Theminimum gaps f are defined between the shaft 1 and the sleeve 2 a at themid portion along the axis. It should be noted that the drawing isseparated by the centerline Z-Z of the bearing axis, thus, the upper andlower parts of the drawing illustrate different configuration of thesleeve 2 a and are combined together for a simple description.

Firstly, the configuration of the sleeve 2 a indicated above thecenterline Z-Z of FIG. 12, will be described hereinafter. The origin Ois defined by the crossing point of the line perpendicular to the axisand extending through the middle point of the sleeve 2 a, and the lineof the outer surface (upper edge) of the shaft 1. Also, the X-Ycoordinate system is defined including the X-axis parallel to thebearing axis and the Y-axis perpendicular to the X-axis. Then, the gap(y-value) of the advantageous configuration of the sleeve 2 a at theposition away from the origin (x-value) is defined by the followingequation;y=k*x ² +f (k: constant)

The advantageous configuration of the sleeve 2 a has a parabolic curvein the drawing. Thus, the above equation can be represented as follows;(y−f)∝x ²

This means that the increased gap α, i.e., (y−f) towards both edges ofthe sleeve 2 a is proportional to the square of the distance from thecenter of the sleeve 2 a.

Another favorable configuration of the sleeve 2 a has an arc curveinstead of the parabola curve. In general, the hydrodynamic bearingassembly has the length L of approximately 20 millimeters, the gap ofseveral microns, and the tilt angle of the shaft relative to the sleeveof approximately 0.05 degrees, which are extremely small values. Thus,the parabola curve can be replaced by a portion of the arc curve havinga great radius. The favorable configuration of the sleeve 2 a is shownin the lower part of FIG. 12. As mentioned above, the upper and lowerparts of the drawing shows different configurations of the sleeve 2 aand 2 a′. The advantageous configuration of the sleeve 2 a′ can bedefined by the arc (a part of circle) having a center on the y-axis andthe radius R expressed by the following formula;R=z*L ²/4/fwherein z has a constant value in range within 0.8 to 1.2.

The present inventors have conducted another experiment for the bearingassembly according to the present invention. The bearing assembly hasthe rotating components including the rotor of 0.1 kilograms weight, androtates at the rotation rate of 20,000 rpm. Also, the shaft has thediameter of 10 millimeters and the bearing assembly has the length of 20millimeters. The bearing assembly was oscillated right and left by 90degrees at the oscillation rate of 90 degrees per 2 seconds. Theoscillation number applied to the bearing assembly was counted until thebearing assembly has a malfunction such as the seizure of the shaft andthe abrasion waste. The result is indicated in Table 3. TABLE 3 Config.Config. Outter Inner Rad. Min. Shaft Sleeve Curve Gap Cnt Oscl. Num. 3-1straight convex  80 m 2 μm YES 200,000 3-2 straight convex 160 m 1 μm NOover 400,000 *3-3  straight straight — 2 μm YES  10,000 *3-4  straightconcave 80 m 2 μm YES  1,000*comparative data

Sample (3-1) has the sleeve 2 a of the convex configuration with theminimum diameter at the mid portion thereof, and the straight shaft.Also, the inner surface of the sleeve 2 a has the radius of 80 meters,and the minimum diameter gap is 2 microns. The hydrodynamic bearingassembly was oscillated around an axis perpendicular to the rotationaxis during the rotation. Then, although contacts between the shaft andsleeve were observed, the bearing assembly of Sample (3-1) workedagainst 200,000 oscillations, and then had a malfunction.

Sample (3-2) is formed similarly to Sample (3-1) except that it has theradius of 160 meters and the minimum diameter gap of 1 micron. Thebearing assembly of Sample (3-2) worked bearing against 400,000oscillations without problems.

Sample (3-3) is the conventional bearing assembly having the straightouter surface of the shaft and the straight inner surface of the sleeve,and also having the minimum diameter gap of 2 microns. The bearingassembly of Sample (3-3) was observed as being worn after 10,000oscillations.

Sample (3-4) of the bearing assembly has the straight outer surface ofthe shaft and the concave inner surface of the sleeve with the maximuminner diameter at the middle portion, and also has the minimum diametergap of 2 microns. The bearing assembly of Sample (3-4) was observed asbeing worn out after 1,000 oscillations.

Fifth Embodiment

The fifth embodiment of the hydrodynamic bearing assembly according tothe present invention, which mainly address improving the bearingrigidity, will be described hereinafter, with reference to drawings.FIG. 13 is a cross section of the hydrodynamic bearing assembly of theembodiment, illustrating sleeve 2 around the shaft 1 a with apredetermined gap F. According to the embodiment, while the sleeve 2 isstraight and has the transverse circular cross section as being evenalong the bearing axis, the shaft 1 a has a barrel configuration, andalso has the transverse cross section with an outer diametercontinuously decreasing towards to the upper and lower ends.

Similar to the fourth embodiment, in the hydrodynamic bearing assemblyof the embodiment, the bearing sleeve 2 has the length L along the axis,the gap between the shaft 1 a and the sleeve 2 is the minimum at theposition of L/2 from the both edges of the sleeve 2. When the shaft 1 aand sleeve 2 are arranged concentrically to each other, the gap betweenthe shaft 1 a and the sleeve 2 is minimized adjacent the middle portion,which is referred to as the one-side minimum gap f. Again similarly, thegap is maximumized adjacent the end portions of the sleeve 2, which isreferred to as the one-side maximum gap fmax. The difference between themaximum gap fmax and the minimum gap f corresponds to the increased gapα. In the embodiment, the position causing the total gap F to be minimumis unnecessarily the exact middle portion, rather it may be adjacent tothe middle portion.

FIG. 14 shows the shaft 1 a and sleeve 2 of FIG. 13 during the rotation.The shaft 1 a is inclined relative to the shaft 2 a because of theexternal biasing force. The gaps become narrower adjacent left upperportion M1 and the right lower portion M2 in FIG. 10, when the spindlemotor includes shaft 1 a rotating inside the sleeve 2. The narrower gapgenerates the higher dynamic pressure, thereby to keep the shaft 1 a andthe sleeve 2 away from each other during the rotation. This is also truefor the spindle motor including sleeve 2 rotating around the shaft 1 a.

The parallel lines indicated adjacent to the end positions M1 and M2 inFIG. 14, schematically illustrate the dynamic pressure distribution. Inorder to minimize the whirl of the shaft 1 a of the motor during therotation, the dynamic pressures generated adjacent to the end positionsM1 and M2 have to be increased across the extensive area by designingthe configuration of the outer surface of the sleeve 2 a so that it hastangential lines substantially parallel to the outer surface of theinclined shaft 1 across the wide area.

The advantageous configuration of the inner surface of the sleeve 2 a ofthe bearing assembly according to the fourth embodiment is applied tothe configuration of the outer surface of the shaft 1 a of the bearingassembly according to the present embodiment. Thus, while the relateddrawing is omitted, the origin O is defined by the crossing point of theline perpendicular to the axis and extending through the middle point ofthe sleeve 2, and the line of the inner surface of the sleeve 2 a. Also,the X-Y coordinate system is defined including the X-axis parallel tothe bearing axis and the Y-axis perpendicular to the X-axis. Then, thegap (y-value) of the advantageous configuration of the shaft 1 a at theposition away from the origin (x-value) is defined by the followingequation;y=k*x ² +f (k: constant)

The advantageous configuration of the shaft 1 a has a parabolic curve inthe drawing. Thus, the above equation can be represented as follows;(y−f)∝x ²

This means that the increased gap α, i.e., (y−f) towards both edges ofthe shaft 1 a is proportional to the square of the distance from thecenter of the shaft 1 a.

The another favorable configuration of the shaft 1 a has an arc curveinstead of the parabola curve. Again similar to the fourth embodiment,the advantageous configuration of the shaft 1 a is convex at the middleportion, and can be defined by the arc (a part of circle) having acenter on the y-axis and the radius R expressed by the followingformula;R=z*L ²/4/fwherein z has a constant value in range within 0.8 to 1.2.

In FIG. 14, the sleeve 2 includes the bevel portions at the endpositions of the sleeve 2 in order to avoid the interference between theshaft 1 a and the sleeve 2, when the shaft 1 a is inclined relative tothe sleeve 2. Alternatively, the convex configuration of the shaft 1 amay slightly be extended towards the sleeve 2, instead of forming thebevel portions.

Sixth Embodiment

The sixth embodiment of the hydrodynamic bearing assembly according tothe present invention, which addresses improving the bearing rigidity,will be described hereinafter, with reference to drawings. In thevertical cross section of the hydrodynamic bearing assembly of thefourth and fifth embodiments, either one of the shaft and the sleeve hasone of the opposing surface of the straight configuration and the otherone of the concave or convex curve so that both surfaces oppose parallelwith gaps during the rotation even when inclined to each other.Meanwhile, the hydrodynamic bearing assembly of the embodiment includesthe shaft and the sleeve both having opposing surface which are convexor concave to realize the similar advantages.

A number of combinations can be conceived for the inner surface of thesleeve and the outer surface of the shaft. The configuration of theinner surface of the sleeve illustrated in FIG. 9 (the fourthembodiment) is referred to as the convex sleeve, and the configurationof the outer surface of the shaft illustrated in FIG. 13 (the fifthembodiment) is referred to as the convex shaft, hereinafter. Then, thecombinations of the curved shaft and the sleeve are; (a) the concaveshaft and convex sleeve, (b) the convex shaft and concave sleeve, and(c) the convex shaft and convex sleeve. The combination of the concaveshaft and concave sleeve is impractical in condition. It should be notedthat in any above combinations, the gaps formed between the shaft andthe sleeve are increased towards the edge portions along the axis in acontinuous and interrupted manner. Also, the gaps are such that theouter surface of the shaft and the inner surface of the sleeve facessubstantially parallel adjacent the end portions (the tangential linesfor both of outer and inner surfaces are substantially parallel as canbe seen in the vertical cross sections), when inclined to each other. Inaddition, preferably, the configurations (curves) of the shaft andsleeve are smooth without any discrete points.

As discussed above, in accordance with the fourth through sixthembodiments, if the shaft and the sleeve are arranged concentrically toeach other, then the one-side maximum gap fmax at the end positions is;fmax=f+α

Thus, the gap is increased by α at the end positions. Since the shaft ismore inclined relative to the sleeve as the increased gap α is greater,the increased gap α is preferably minimized. According to theembodiment, the increased gap α is preferably about 2 microns or less,and more preferably about 1 micron or less, however, if it is about 0.1micron or less, then the aforementioned effects can not be achievedbecause it is too small.

The fourth through sixth embodiments have been discussed for mainlyimproving the bearing rigidity by inclining to rotate either one theshaft and the sleeve relative to the other towards the predetermineddirection. A device for biasing and inclining either one of the shaftand the sleeve relative to the other towards the predetermined directionwill be described hereinafter. FIG. 16 illustrates one example of thebiasing device. In FIG. 16, the bearing assembly comprises the shaft 31and the sleeve 32 arranged around the shaft, in which the sleeve 32 hasthe inner surface diameter increasing towards the ends of the sleeve 32.The shaft 31 has the rotor 33 secured thereon. The rotor 33 has a skirtwith an inner surface provided with a rotor magnet 34, which opposes tothe stationary stator 35 secured on the sleeve 32. A ring-shape magnet36 is attached on the top surface of the sleeve 32, which is eccentricto the axis of the sleeve 32 and opposes to an another ring-shapedmagnet 37 provided around the shaft 31. Both magnets 36 and 37 arearranged so that the repulsion force is generated between them.

In the spindle motor so constructed, when a coil (not shown) wound onthe stator 35 is energized by the electric flow, to thereby generate theattraction force (or the repulsion force), the rotor 33 rotates with theshaft 31 relative to the sleeve 32. The rotor 33 is kept away from thesleeve 32 due to the repulsion force between two ring-shaped magnets 36and 37. As described above, the ring-shaped magnet 36 is eccentric tothe ring-shaped magnet 37, when the repulsion force is the greatest atthe position where two ring-shaped magnets 36 and 37 are closest. Thisforce biases and inclines the shaft in the direction from the closestposition to the farthest position. In the drawing, the reference numeral38 denotes the central axis of the sleeve 32 and the reference numeral39 denotes the rotation axis of the sleeve 32.

As discussed above, when the shaft 31 is inclined and rotated relativeto the sleeve 32, the dynamic pressure generated between the shaft 31and the sleeve 32 adjacent the portions where the both are closer toeach other. The dynamic pressure is such force that the sleeve 32 pushesthe shaft 31 away. Therefore, the repulsion force by the magnets 36, 37and the dynamic pressure together balance the sleeve 32 and the shaft31, thereby to realize the relative rotation in a stable manner.

In FIG. 16, the ring-shaped magnets 36 and 37 are eccentricallypositioned on the top surface of the bearing assembly. However, thoseeccentric magnets may be provided on the bottom surface of the bearingassembly at such portion that the repulsion force by the magnets biasesthe shaft in the same direction. Also, two pairs of eccentric magnetsmay be provided on both top and bottom surfaces of the bearing assemblyas well for inclining the shaft in a even more stable manner.

Although FIG. 16 illustrates the shaft 31 as being the rotational memberand the sleeve 32 as being the stationary member, they may be changedwith each other. In this instance, the eccentric magnet has to beprovided on the stationary member. Also, the attraction force, insteadof the repulsion force, between the magnets can be used to bias theshaft to the predetermined direction. In this case, one of the magnetmay be substituted by a ferromagnet.

Seventh Embodiment

The seventh embodiment of the hydrodynamic bearing assembly according tothe present invention, which mainly addresses improving the bearingrigidity, will be described hereinafter, with reference to drawings.FIG. 17 illustrates one example of the spindle motor incorporating thehydrodynamic bearing assembly of the embodiment. In the drawing, adisk-shaped thrust plate 43 is arranged on a portion of a shaft 41,extending in a plane perpendicular to the bearing axis. The shaft 41 andthe thrust plate 43 are received within a sleeve 42, with apredetermined gap therebetween. A plurality of grooves for generatingthe dynamic pressure are formed either on opposing surfaces of the shaft43 and the sleeve 42. In other words, the hydrodynamic bearing assemblyhas the thrust bearing between a pair of the radial bearings. A rotor 44secured on the shaft 41 has a skirt with the inner surface provided witha rotor magnet 45. The rotor magnet 45 opposes to a stator 46 secured onthe sleeve 42.

In the operation of the spindle motor so constructed, the electriccurrent provided with the stator 46 generates the attraction and/or therepulsion forces between the stator 46 and the rotor magnet 45 togenerate the rotational driving force of the rotor 44 having the rotormagnet 45, thereby to rotate the rotor 44 together with the shaft 41.This rotation defines the hydrodynamic bearings, in which the shaft 41and the thrust plate 43 rotate relative to the sleeve with thepredetermined gap therebetween.

According to the embodiment, a thrust tilt angle is defined as an anglewhen the opposing surfaces of the thrust plate 43 and the sleeve 42 arein contact with each other, also a radial tilt angle is defined as anangle when the outer surface of the shaft 41 and the inner surface ofthe sleeve 42 are in contact with each other. The hydrodynamic bearingassembly according to the embodiment has the thrust tilt angle greaterthan the radial tilt angle to prevent the thrust plate 43 fromcontacting with the sleeve 42.

FIG. 18 is an enlarged view of the bearing. Similar components aredenoted by similar reference numerals. The inclined shaft 41 generatesthe dynamic pressure distributions in the radial and thrust bearings asschematically illustrated by the parallel lines M1, M2 and N1 N2,respectively, in the drawing.

When the shaft 41 is inclined relative to the sleeve 42, the narrowergap between the shaft 41 and the sleeve 43 generates the higher dynamicpressure adjacent the portions denoted by M1 and M2 in the radialbearing. The higher dynamic pressure bears the shaft 41 against furthertilt to keep it away from the sleeve 42 (non-contact). Meanwhile, sincethe hydrodynamic bearing assembly is one of the pump-in type and thefluid is conducted from the circumference of the trust plate 43, thedynamic pressure distribution cannot have a peak, rather a relative lowdynamic pressure at the circumference. Thus, when the thrust plate 43 isinclined and the opposing surfaces of the sleeve 42 is close the thrustplate 43, the moment generated is unlikely enough to bear against theexternal motion.

Referring back to FIG. 17, several dimensions in the drawing will bedefined hereinafter. The radial bearing has the length of L. The shaft41 has the right-side gap f1 and left-side gap f2 relative to the sleeve42 as well as the total gap F (F=f1+f2). The trust bearing has the outerdiameter G. The thrust plate 43 has the upper gap d1 and the lower gapd2 relative to the opposing surface of the sleeve 42 as well as thetotal gap D (D=d1+d2). When the external motion is applied so that theshaft is inclined relative to the sleeve, the radial bearing has thedynamic pressure sufficient to bear the shaft against the further tiltprior to the contact between the thrust plate 43 and the opposingsurface of the sleeve 42 for keeping them away from each other.

Should there be no thrust plate 43, then the shaft 41 could be inclinedrelative to the sleeve 42 at the greatest angle of F/L. On the otherhand, should there be no shaft 41, then the thrust plate 42 could beinclined relative to the sleeve 42 at the greatest angle of D/G. Asmentioned above, in order for the radial bearing to have the dynamicpressure great enough to prevent the contact between the thrust plate 42and the opposing surface of the sleeve 42, the greatest angle F/L in theradial bearing is to be less than the greatest angle D/G in the thrustbearing. Thus, the contact in the thrust bearing is avoided, when thefollowing condition is satisfied.(F/L)<(D/G)In case where the shaft 41 and sleeve 42 may be processed to have somecircularity deviation and cylindricality, the dimensions F and D areregarded as the average length and the diameter, respectively.

In the embodiment, the tilt of the shaft 41 (and/or the thrust plate 43)relative to the shaft 42 has been discussed. The dynamic pressure isunlikely generated in the thrust bearing for bearing against theexternal motion to return to the normal position when the thrust plate43 is inclined to the opposing surface of the sleeve 42. However, whenthe axis of the shaft 41 is maintained parallel to the bearing axis, thethrust plate 43, as a whole, generates a force enough for supporting theshaft 41. To this end, the gap in the radial bearing is much narrowerthan the gap in the thrust bearing so that the thrust plate 43vertically shifts in parallel with the opposing surface of the sleeve 42to prevent the thrust plate 43 from contacting with the sleeve 42.

According to the prior arts, while the radial gap along the diameter is6 to 10 microns (the on-side radial gap along the radius is 3 to 5microns), the thrust gap is along the bearing axis is 4 to 6 microns,thus, the ratio thereof is about 1. However, as the details will bedescribed later, the experiments that the present inventors haveconducted revealed that those radial and thrust gaps are insufficient,they are required to satisfy at least the following condition;2F<DPreferably,4F<DTherefore, in general, the radial gap F and the thrust gap D arerequired to have following relation to avoid the contact in the thrustbearing in the embodiment;kF<D (k is constant in the range of 2 to 10)

Although discussions have been made above for the hydrodynamic bearingassembly illustrated in FIG. 17, in which the thrust bearing is arrangedbetween the pair of radial bearings, the present invention can also beapplied to the hydrodynamic bearing assembly illustrated in FIG. 19, inwhich a pair of the thrust bearing 43 a and 43 b sandwich the sleeve42′. In FIG. 19, the pair of the thrust bearings 43 a and 43 b issecured to the shaft 41′, the sleeve 42′ is arranged between the thrustbearings to define a pair of the thrust bearings adjacent to the upperand lower ends of the sleeve 42′. The thrust bearings have the upper gapd1 and the lower gap d2 relative to the opposing surfaces of the sleeve42 as well as the total gap D (D=d1+d2). Also, the radial bearing hasthe right-side gap f1 and the left-side gap f2 as well as the total gapF (F=f1+f2). Then, the contact in the thrust bearing is avoided, if thethrust tilt (gradient) of D/G is less than the radial thrust tilt(gradient) of F/L, i.e. the following condition is met;(F/L)<(D/G)Further, preferably, the following condition is met;kF<D (k is constant in the range of 2 to 10)

In general, a hydrodynamic bearing assembly, in which a radial bearingis arranged between a pair of the thrust bearings, or a thrust bearingis arranged between a pair of the radial bearings, have the thrust gapsatisfying the above-mentioned conditions to prevent the contact in thethrust bearing.

Alternatively, although the aforementioned disclosure is made for therelation between the radial gap and the thrust gap, the extremely narrowradial gap causes the thrust plate 43 without meeting the aboveconditions to be kept away substantially in parallel with the opposingsurface of the sleeve. For example, the total radial gap F in thediameter direction not greater than about 6 to 10 microns (about 3 to 5microns on each side) alone prevents the contact in the thrust bearingdue to the tilt of the shaft, and provides the supporting force forbearing the vertical motion. Therefore, the alternative approach toimprove the bearing rigidity is to set the total radial gap F of about 3microns or less, and preferably about 2 microns or less. Theoretically,although the total radial gap F might be designed to be zero, the sleevecould not be rotated relative to, or even be arranged around the shaftwhen the processing accuracy of the outer surface of the shaft 41 andthe inner surface of the sleeve 42 are not well controlled. Therefore,in practical, the total radial gap F is designed to be 1 micron or morefor achieve the improvement of the bearing rigidity. In addition,reduction of the radial gap F can also eliminate the half-whirlphenomenon as described in the first embodiment.

The present inventors have conducted an evaluation test for the spindlemotor of FIG. 17 incorporating the hydrodynamic bearing assemblyaccording to the embodiment. The rotational member including the rotorhave the mass of 0.1 kilograms, the rotation rate of 12,000 rpm. Theradial bearing has the diameter of 10 millimeters and the length of 20millimeters, and the thrust bearing has the diameter of 20 millimeters.The radial gap and the thrust gap are designed to be 5 microns and 1micron, respectively. When the external motion by twisting with hand wasapplied to the bearing axis of the spindle motor around the axisperpendicular to the bearing axis, no sound showing the contact in thebearing assembly was observed. Contrary, when the similar externaloscillation by twisting with hand was applied to the bearing axis of thespindle motor having the thrust gap of 5 microns and the radial gap of 3microns, the sound showing the contact in the bearing assembly was infact heard.

The third aspect according to the present invention mainly addressesimproving the bearing rigidity, and in particular, relates to thepump-out type hydrodynamic bearing assembly. As discussed above,according to the pump-out type hydrodynamic bearing assembly, since thethrust dynamic pressure distribution can be generated to have thedynamic pressure peak adjacent the thrust plate, the tilt rigidityagainst the external oscillation force is advantageously improved.However, according to the conventional bearing assembly, there is adisadvantage that the fluid is insufficiently provided to the pump-outthrust bearing. The pump-out type hydrodynamic bearing assemblyaccording to the embodiments of the present invention overcomes thedisadvantage.

Eighth Embodiment

The eighth embodiment of the hydrodynamic bearing assembly according tothe present invention will be described hereinafter, with reference todrawings. FIG. 20 shows the hydrodynamic bearing assembly of theembodiment. The hydrodynamic bearing assembly comprises a shaft 1, asleeve 2 rotatably arranged around the shaft 1, and a disk-shaped thrustplate 3 perpendicularly secured onto the shaft 1. One end surface of thesleeve 2 (referred to as a thrust opposing surface 4) opposes to thethrust plate 3. A thrust bearing is defined between the thrust plate 3and the thrust opposing surface 4. A plurality of spiral grooves 5 isformed on the thrust plate 3 for generating the pump-out type dynamicpressure. Also, a radial bearing is defined between the outer surface ofthe shaft 1 and the inner surface of the sleeve 2. Although FIG. 20illustrates the shaft 1 having one longitudinal groove 9 formed on theouter surface thereof, preferably it includes a plurality oflongitudinal grooves. In this drawing, the stationary member include theshaft 1 and the thrust plate 3, and the rotational member include thesleeves and other components (not shown).

In the thrust bearing, since the thrust plate 3 includes the pomp-outtype spiral grooves 5, the fluid is conducted from the portions adjacentto the bearing axis to the circumference of the thrust plate 3 asindicated by arrow 24, due to the relative rotation of the thrustbearing. Because the conducted flow of the fluid generates high dynamicpressure adjacent the circumference of the thrust plate 3, the dynamicpressure distribution in the thrust bearing is obtained as indicated bythe dashed line 25. The dynamic pressure with its peak adjacent to thecircumference causes the thrust bearing to be improved in the rigidityagainst the external oscillating moment.

Because of the radial dynamic pressure, the sleeve 2 rotates relative tothe shaft 1 without contact in the radial bearing. The fluid issequentially supplied via the longitudinal groove 9 formed on the shaft1 to the direction indicated by the arrow 26 so that the thrust bearingreceives the fluid enough to generate the good thrust dynamic pressures.In addition, the longitudinal grooves 9 contributes in preventing theabove-mentioned half-whirl. When the depth of the longitudinal groove 9is so deep, the dynamic pressures in the radial bearing and the thrustbearing are close to the atmosphere to lose the radial rigidity and toreduce the thrust dynamic pressure. The present inventors have foundthat the depth ratio (which is referred to as a ratio of the depth ofthe longitudinal groove 9 relative to the effective diameter of theshaft 1) is advantageously set to 0.005 or less, and even morepreferably set to 0.001 or less. However, when the depth of thelongitudinal groove 9 is too shallow, the flow of the fluid can beinsufficiently provided. Thus, the aforementioned depth ratio ispreferably 0.0001 or more.

The disadvantage caused by the longitudinal groove 9 indicated in FIG.20, in general, results the dynamic pressure to be reduced locallyadjacent to the longitudinal groove 9, and hence the translationrigidity to be reduced in the direction of the longitudinal groove 9.One approach to avoid the disadvantage is to provide the longitudinalgrooves 9 on the inner surface of the sleeve 2 instead of the outersurface of the shaft 1. Since the rotational member generally rotate atthe rotation rate of 10,000 rpm or higher, the dependency upon thedirection of the grooves can be disregarded.

Another approach to avoid the disadvantage is to make the longitudinalgrooves 19 inclined relative to the bearing axis as shown in FIG. 21.The inclined grooves 19 cause the portions (portions adjacent to thegroove 19) having the low dynamic pressure to be helical surrounding onthe outer surface. This generates the dynamic pressure that bearsagainst the external force in any directions, thereby to reduce thepossibility of the contact. The inclined grooves 19 provide anotheradvantage that the rotation of the sleeve 2 in the direction asindicated by the arrow 28 propels the fluid due to its viscosity alongthe direction indicated by the arrow 28. This allows the fluidsufficiently provided into the thrust bearing for increasing the dynamicpressure therein. The bearing assembly with the inclined grooves 19 canbe designed by selecting the numbers of grooves and the varying theinclined angle relative to the bearing axis so that the fluid can beprovided into the thrust bearing as desired.

Even another approach for propelling the fluid into the thrust bearingand for minimizing the reduction of the translation rigidity is to forma plurality of intermittent inclined grooves 19 a and 19 b as indicatedin FIG. 22. As illustrated in the drawing, a plurality of groovesinclined relative to the bearing axis are formed on the outer surface,which are intermittent adjacent the middle portion of the shaft 1 toform the upper groove 19 a and the lower groove 19 b. When the outersurface of the shaft 1 and the inner surface of the sleeve 2 have nocircularity deviation, the flow of the fluid is inadequately developed.However, if the grooves provided for propelling or for drawing the fluidin cooperation with a portion with no circularity deviation is locatedin the mid portion or end of the shaft 1, a predetermined amount of thefluid in the direction of the bearing axis can be assured to providewith the grooves. In the drawing, the upper inclined groove 19 a propelsthe fluid down indicated by the arrow 29 due to the rotation of thesleeve, and the lower inclined groove 19 b similarly draws the fluiddown. This allows the flow of the fluid to be promoted down towards thebearing axis in comparison with the bearing assembly with no circularitydeviation. The flow of the fluid so provided can be effectively used inthe pump-out thrust bearing to generate much dynamic pressure therein.Since the fluid flown through the radial gap between the shaft 1 and thesleeve 2 is squeezed in the mid portion of the bearing assembly, therestriction conditions are determined based upon the squeezed portion.This allows the various dimensional requirements in accuracy to be lessprecisely so that the various dimensions are selected relatively in aflexible manner. This is another advantage due to the intermittentgrooves 19 a and 19 b. The dependency of the direction of the externalmotion relative to the shaft due to the grooves is eliminated by theportion without circularity deviation, which is an another advantage.

The embodiment has been described that the hydrodynamic bearing assemblyhas the grooves formed on the outer surface of the shaft for conductingthe fluid, such grooves are formed on the inner surface. Also, in casewhere the grooves are inclined to the bearing axis, preferably, theinclined grooves are such that the fluid is propelled down to increasethe dynamic pressure in the thrust bearing. The shaft 21 may be eitherrotational member and the stationary member. Alternatively, the spiralgrooves 5 for generating the thrust dynamic pressure may be formed onthe thrust opposing surface of the sleeve 2. Although the FIG. 22illustrates three inclined grooves divided into upper and lower grooves19 a and 19 b, they may be divided into three or more parts by the twoor more intermittent portions.

Ninth Embodiment

The ninth embodiment of the hydrodynamic bearing assembly according tothe present invention, which mainly addresses improving the bearingrigidity by incorporating the pump-out type hydrodynamic bearingassembly, will be described hereinafter, with reference to drawings.FIG. 23 illustrates one example of the spindle motor incorporating thehydrodynamic bearing assembly of the embodiment. In the drawing, a shaft51 is secured onto a base plate 50 having the column members 51 a and 51b with relatively small diameter. A sleeve 52 having hollow cylindricalinner surface parallel to the outer surface of the shaft 51 is rotatablyarranged around the shaft 51. The sleeve 52 includes a first and seconddoughnut thrust plates 55 a and 55 b both sandwiching the sleeve 52. Thedoughnut thrust plates 55 a and 55 b have the upper and lowerthrough-holes in the mid portion thereof through which the columnmembers 51 a and 51 b are extending. The end surface of the sleeve 52opposing to the first thrust plate 55 a is referred to as a firstopposite surface 54 a, and the another end surface of the sleeve 52opposing to the second thrust plate 55 b is referred to as a secondopposite surface 54 b. Either one of the first and second oppositesurfaces may be integrally formed with sleeve 52.

The thrust bearings are defined between the first thrust plate 53 a andthe first thrust opposing surface 54 a, also between the second thrustplate 53 b and the second thrust opposing surface 54 b. Also, the radialbearing is defined between the outer surface of the shaft 51 and theinner surface of the sleeve 52. The rotor 57 is secured onto the outersurface of the sleeve 52 and have a skirt provided with a rotor magnet58 opposing to the electromagnet 59 secured on the rotor 57. Theelectromagnet 59 energized by the electric flow generates the attractionforce (or the repulsion force), the sleeve 52, the rotor 57 and otherrotational member rotate relative to the shaft 51 in a similar manner asthe other hydrodynamic bearing assemblies.

In operation of the spindle motor as described above, the flow of thefluid and the dynamic pressure distributions in the radial and thrustbearings will be described hereinafter, with reference to FIG. 24. FIG.24 is an enlarged view of the hydrodynamic bearing assembly of FIG. 23,and similar parts are denoted by the similar reference numerals. In thedrawing, a plurality of the pump-out type spiral grooves 56 a are formedon the first thrust plate 53 a, and a plurality of the pump-in typespiral grooves 56 b are formed on the second thrust plate 53 b. Thosespiral grooves 56 a and 56 b may be formed on the thrust opposingsurfaces 54 a and 54 b, instead of on the first and second thrust plates53 a and 53 b.

The rotation of the sleeve 52 around the shaft 51 causes the relativerotation between the first thrust plate 53 a and the first oppositesurface 54 a. The groove 56 a in cooperation with the relative rotationconducts the fluid in a direction indicated by the arrow 61 through thegap between the small column 51 a and the first thrust plate 53 a intothe thrust bearing assembly. The pump-out type spiral grooves 56 acontinuously propels the fluid to the circumference of the first thrustplate 53 a to increase the dynamic pressure towards the circumferenceaway from the bearing axis, thereby to generate the thrust dynamicpressure for bearing the first thrust plate upwardly. Then, the fluid isguided in the direction indicated by the arrow 62 to the radial bearingconnecting with the thrust bearing, and generates the radial dynamicpressure between the outer surface of the shaft and the inner surface ofthe sleeve 52. The fluid carried to the bottom surface of the shaft 51is conducted again into the other thrust bearing. The pump-in typespiral grooves 56 b in cooperation with the relative rotation betweenthe second thrust plate 53 b and the second opposite surface 54 b drawsthe fluid from the circumference of the second thrust plate 53 b towardsthe bearing axis. The fluid is compressed by the grooves 56 b adjacentto the axis, thereby to generate the thrust dynamic pressure bearing thesecond thrust plate downwardly. Then, the fluid is expelled outwardlyvia the gap between the second thrust plate 53 b and the through-hole 55b as indicated by the arrow 63.

FIG. 24 also indicates the parallel lines schematically representatingthe dynamic pressure distributions in the first and second thrustbearings adjacent to the portions indicated by the parallel lines. Ascan be seen in the dynamic pressure distributions, the pump-out typespiral grooves 56 a generates the peak of the dynamic pressure adjacentto the circumference away from the bearing axis for bearing the shaft 51against the external oscillation to prevent the shaft to tilt. On theother hand, the pump-in type spiral grooves 56 b generates the peak ofthe dynamic pressure adjacent to the bearing axis for bearing the secondthrust plate 53 b and the sleeve 52 downwardly as described above. Inother words, the dynamic pressure in the second thrust bearing causes adepressing force to the sleeve 52. The dynamic pressure is, in general,increased as the members rotating relative to each other have a closergap therebetween. The depressing force to the sleeve 52 also depressesthe opposite first thrust plate 53 a downwardly to cause the thrustdynamic pressure generated by the pump-out type spiral grooves 56 a tobe increased. As described above, according to the hydrodynamic bearingassembly of the embodiment, a pair of thrust bearings provided on theupper and lower end portions of the sleeve 51 together cooperate toavoid the tilt against the external motions, thereby to achieve theanti-tilt hydrodynamic bearing assembly. Unlike the conventionalpump-out type hydrodynamic bearing assembly indicated in FIGS. 55 and56, since the thrust bearings according to the embodiments are open tothe atmosphere adjacent to the bearing axes, a sufficient amount of thefluid can be supplied into the thrust bearings, thereby to effectivelygenerate dynamic pressure enough to prevent the tilt.

It should be noted that the similar advantage is enjoyed by conductingthe fluid flow in the reverse direction to one indicated in FIG. 24. Thefirst and second thrust bearings are designed as the pump-in andpump-out type ones, respectively, so that the fluid is conducted fromthe bottom and expelled towards the top. This achieves the hydrodynamicbearing assembly with the robust rigidity similar to theabove-mentioned.

FIGS. 25(a) and 25(b) illustrate the hydrodynamic bearing assemblies ofthe embodiment, further including longitudinal grooves 64, 64 a formedon the outer surface of the shaft 51, which are parallel and inclined,respectively, to the bearing axis. Those longitudinal grooves 64, 64 apromote the sequential fluid flow from the pump-out inlet thrust bearingto the pump-in outlet thrust bearing. To this end, the sufficient amountof the fluid is smoothly supplied to increase the dynamic pressure. Inparticular, the inclined longitudinal grooves 64 a of FIG. 25(b) incooperation with the counterclockwise rotation (as seen from the top)causes the fluid to smoothly flow and increases the dynamic pressuremore effectively.

The advantages and disadvantages are as discussed above for the eighthembodiment. The longitudinal grooves are preferably inclined to the axisespecially for minimizing the reduced translation rigidity due to thegrooves. Also, the total number of the grooves formed on the shaft 51 ispreferably in the range within about three to six. The gradient of thegroove may be expressed by a central angle defined the line from thecenter of the axis to start point of the groove and the line from thecenter of the axis to end point of the groove. Even if the central angleis only about 10 degrees, the advantage of the promotion of the fluidsupply can be obtained to some extent. However, the central angle ispreferably 30 degrees or more to avoid the reduced translation rigidity.When the shaft 22 a has three grooves 64 a formed thereon, each centralangle of the grooves preferably is 120 degrees so that the wholecircumference of the shaft 22 a has the grooves 64 a. However, thegrooves 64 a are not necessarily formed on the whole circumference, morepreferably, the central angle of the grooves falls within the range of30 to 60 degrees.

Although the grooves 64 and 64 a are formed on the outer surface of theshaft 51 according to the embodiment, they may be formed on the innersurfaces of the sleeve 52 to achieve the similar advantages. Also, theshaft may be designed as the rotational member instead of the stationarymember similar to the other embodiments as discussed above.

FIG. 26 illustrates another example of the hydrodynamic bearing assemblyincluding the upper (first) and lower (second) thrust bearings, whichare both pump-out type ones. In FIGS. 26(a) to 26(d), the pump-out typespiral grooves 56 c for generating the dynamic pressure are formed onthe second thrust plate 53 c as the first thrust plate 53 a, whichopposes to the second thrust opposing surface 54 b of the shaft 51. Theother structure of the bearing assembly is similar to those in FIG. 23.

According to the hydrodynamic bearing assembly of FIG. 26(a), upon therotation of the sleeve 52 relative to and around the shaft 51, thepump-out type first thrust bearing between the first thrust plate 53 aand the first thrust opposing surface 54 a conducts the fluid into thebearing assembly as indicated by the arrow 61. Also, the pump-out typesecond thrust bearing between the second thrust plate 63 a and thesecond thrust opposing surface 54 b conducts the fluid into the bearingassembly as indicated by the arrow 67.

The fluid conducted by both thrust bearing assemblies are propelled fromthe central portions to the radially outer portions by the pump-out typespiral grooves 56 a and 56 c, so as to generate the thrust dynamicpressures with peaks on the circumferences, as indicated by the parallellines on the first thrust plate 54 a of FIG. 24. In the hydrodynamicbearing assembly of FIG. 26, since both thrust bearings are the pump-outtype ones, the thrust dynamic pressures with peaks on the circumferencesare generated, thus, to achieve the hydrodynamic bearing assembly withthe robust rigidity against the tilt-motion. In addition, since thefluid are conducted through both ends to increase the dynamic pressurein the radial bearing, thereby increasing the bearing force in theradial bearing.

The fluid conducted into the hydrodynamic bearing assembly is expelledby the dynamic pressure continuously supplied in the bearing assembly,through the through-holes 55 a and 55 c between the small columns 51 aand 51 b of the shaft 51 and the thrust bearing plate 53 a and 53 c tothe atmosphere. The spiral grooves 56 a and 56 c for generating thethrust dynamic pressure may be formed on the thrust opposing surfaces 54a and 54 b instead of on the thrust plate 53 a and 53 c.

According to the hydrodynamic bearing assembly of FIG. 26(b), thelongitudinal grooves 64 are formed on the outer surface of the shaft 51extending parallel to the bearing axis between the first and secondthrust bearings. The longitudinal grooves effectively communicate influid between the first and second thrust bearings so that both thrustbearings have the even dynamic pressure. The even dynamic pressurecauses the rotational member such as the sleeve 52 to be kept in astable axial position, and prevents the half-whirl phenomenon. Also thelongitudinal grooves may be extended from the both ends but interruptedin the mid portion of the radial bearing to form upper and lower groovesnot communicating each other. This structure can further increases thedynamic pressure adjacent the middle portion of the radial bearing toimprove the bearing rigidity in the radial bearing.

According to the hydrodynamic bearing assembly of FIG. 26(c), aplurality of herringbone grooves 68 is formed on the outer surface ofthe shaft 51. The herringbone grooves 68 have gradients such that thefluid is guided towards the middle portion of the radial bearing uponthe relative rotation of the shaft 51 and the sleeve 52. In FIG. 26(c),for example, the fluid is guided towards the mid portion of the sleeve52 as indicated by the arrows 69 and 71 when the sleeve 52 rotatescounterclockwise (as seen from the top) as indicated by the arrow 66.Therefore, the herringbone grooves 68 increases the dynamic pressureadjacent the middle portion of the radial bearing, thereby to improvethe bearing rigidity in the radial bearing. Thus, the herringbonegrooves 68 can compensate the shorter radial bearing along the axis, forgenerating the sufficient radial dynamic pressure. Also the herringbonegrooves 68 uniformalizes the dynamic pressures in both thrust bearingsor eliminates the half-whirl. Various configurations of the herringbonegrooves 68 can be conceived rather than the V-shaped one, including theV-shaped grooves interrupted in the mid portion of the radial bearing,the V-shaped grooves emerged into an annular groove in the mid portionof the radial bearing, and the V-shaped grooves but asymmetricallyformed. The herringbone grooves 68 may have any configuration as long asthey can conduct the fluid from both thrust bearings towards the middleportion of the radial bearing.

According to the hydrodynamic bearing assembly of FIG. 26(d),through-holes are provided inside the shaft 51 for expelling the fluidconducted from both of the pump-out type thrust bearings. In thehydrodynamic bearing assembly of FIG. 26(a), the fluid is introduced andexpelled through the same gaps defined between the small columns 51 a,51 b and the through-holes 55 a, 55 b of the first and second thrustplates 53 a, 53 b. In some cases where the viscosity of the fluid ishigh, the conflict between the introduced fluid and the expelled fluidprevents the fluid from effectively introducing into the bearingassembly by means of the pump-out type spiral grooves. To address thedisadvantage, the bearing assembly of FIG. 26(d) has another passage forexpelling the fluid, as well as the gap defined between the shaft 51 andthe through-holes 55 a, 55 b of the thrust plates 53 a, 53 b forintroducing the fluid into the bearing assembly.

As illustrated in FIG. 26(d), a longitudinal hole is provided within theshaft 51, extending through or adjacent to the bearing axis. Also,preferably, a plurality of transverse holes 73 are provided within theshaft 51, extending perpendicularly from the outer surface of the shaft51 adjacent the middle portion of the radial bearing to the longitudinalhole for fluid communication between the longitudinal hole 72 and thetransverse holes 73. The bearing assembly of FIG. 26(d) has thelongitudinal hole 72 extending through the first and second endsurfaces, however, the longitudinal hole 72 may extend through eitherone of upper and lower surfaces if it is connected with the transverseholes 73. This structure allows the fluid introduced by the pomp-outtype spiral grooves 56 a and 56 c of the thrust bearings to generate thethrust and radial dynamic pressures in the thrust and radial bearings,respectively, and then to be guided into the transverse holes 73 to beexpelled via the longitudinal hole 72 to the atmosphere. Thelongitudinal hole 72 and transverse holes 73 may be provided within theshaft in cooperation with the grooves of FIGS. 26(b) and 26(c). Althoughthe holes for expelling the fluid reduce the dynamic pressure locally inthe radial bearing, if the radial dynamic pressure can be assured, forexample, by having the sufficient long radial bearing, the fluid flowcan be promoted to increase the dynamic pressure especially in thethrust bearings. If feasible for use, the holes for expelling the fluidmay be provided adjacent the middle portion of the sleeve 52, instead ofthe shaft 51, outwardly extending to the atmosphere.

FIG. 27 illustrates the hydrodynamic bearing assembly similar to one ofFIG. 23 except that one of the thrust bearing has a closed end coveredby the thrust plate of the sleeve 52. A column-shaped shaft 81 has asmall column formed concentrically with the shaft 81, also has an outersurface parallel to the axis of the shaft 81. A cylindrical hollowsleeve 52 is rotatably arranged around the shaft 81. At one end, thesleeve 52 is secured to a first thrust plate 83 a having a through-hole85, through which the small column 81 a of he shaft 81 is extending. Thefirst thrust plate 83 a opposes to the bottom surface of the shaft 81,referred to as the first opposite surface. At the other end, the sleeve52 is secured to a disk-shaped second thrust plate 83 b covering thehollow portion of the sleeve 52. The second thrust plate 83 b opposes tothe upper surface of the shaft 81, referred to as the second oppositesurface. The second thrust plate 83 b has no through hole for the smallcolumn. Also, the first and second thrust plate 83 a and 83 b have thepump-out type and pump-in type spiral grooves 86 a and 86 b, and opposethe thrust opposite plates 84 a and 84 b, respectively. Further, eitherone of the first and second thrust plates 83 a and 83 b may beintegrally formed with the sleeve 52. The spiral grooves 86 a and 86 bmay be formed on the thrust opposing surfaces 84 a and 84 b.

Upon the relative rotation between the shaft 81 and the sleeve 52, thefluid introduced by the pump-out type spiral grooves 86 a as indicatedby the arrow 87 generates the thrust dynamic pressure between the firstthrust plate 83 a and the first opposite surface 84 a. Then, the fluidis conducted into the radial bearing to generate the radial bearing. Thefluid is guided as indicated by the arrow 88 and compressed as indicatedby the arrow 89 into a second thrust bearing between the second thrustplate 83 b and the second opposite surface 84 b, thereby to generatethrust dynamic pressure in the second thrust bearing by means of thespiral grooves 86. Then, the dynamic pressure generated by thecontinuous fluid flow expels the fluid out to the atmosphere by throughthe gap between the small column 81 a of the shaft 81 and thethrough-hole 85 of the first thrust plate 83 a, accordingly.

The hydrodynamic bearing assembly of FIG. 27 also generates the peak ofthe dynamic pressure on the circumference away from the bearing axis inthe thrust bearing, thereby to bear against the tilt moment. Thus, thebearing assembly has the improved robust rigidity. The bearing assemblyof FIG. 27 also has the inclined grooves 91 formed on the outer surfaceof the shaft 81 for guiding the fluid from the pump-out type thrustbearing to the pump-in type thrust bearing. The grooves may be straightto the bearing axis or the herringbone grooves. Alternatively, no groovemay be formed on the shaft 81. In the hydrodynamic bearing assemblyaccording to the embodiment, the gap between the small column 81 a ofthe shaft 81 and the through-hole 85 of the first thrust plate 83 a isused both for introducing and expelling the fluid, this might cause theadverse effect to the fluid flow. In order to avoid such adverse effect,the hydrodynamic bearing assembly may have another hole indicated by thedashed line 92 in the second thrust plate 83 a, or a through-holeextending through the axis of the shaft 81 similar to one illustrated inFIG. 26(d)

FIG. 28 illustrates one example of a shaft-rotating spindle motorincorporating the hydrodynamic-bearing assembly of FIG. 27. In thedrawing, the sleeve 52 including the first and second thrust bearings 83a and 83 b is supported on the base plate 80, and the shaft 81 havingthe small column 81 a formed concentrically with the shaft 81 isrotatably arranged within the inner surface of the sleeve 52. Thisstructure defines the radial bearing and a pair of the thrust bearingscommunicating with the radial bearing at the both ends of the shaft 81.The rotor 93 is secured onto the column 81 a of the shaft 81, and hasthe rotor magnet 94 in the inner surface thereof. The rotor magnet 94opposes to the electromagnet 95 secured on the base plate 80. Theelectromagnet 95 energized by the electric current generates theattraction force or the repulsion force between the electromagnet 74 andthe rotor magnet 73, thereby to rotate rotational member such as theshaft 81 and rotor 93 around the sleeve 52. The rotation of the spindlemotor is the same as the others as described above.

Since the peak of the dynamic pressure can be generated at thecircumference away from the bearing axis, the spindle motor as describedabove has the improved robust rigidity.

The fourth aspect according to the present invention relates to thehydrodynamic bearing assembly in which the shaft is securedperpendicularly onto the thrust plate with the high accuracy in a simplemanner in order to improve the bearing rigidity. As discussed above, itis critical to keep the perpendicularity between the shaft of the radialbearing and the thrust plate of the thrust bearing with a high accuracy,so as to avoid the contact of the thrust bearing especially during therotation.

Tenth Embodiment

The tenth embodiment of the hydrodynamic bearing assembly according tothe present invention, which addresses improving the bearing rigidity byassuring the accuracy of the perpendicularity between the shaft and thethrust plate, will be described hereinafter, with reference to drawings.FIG. 29 is a cross section of the hydrodynamic bearing assembly of theembodiment. In the drawing, a hollow cylindrical shaft 101 isperpendicularly positioned on a disk-shaped thrust plate 103 with athrough-hole 104. As described above, if the perpendicularity betweenthe members 101 and 103 has insufficient accuracy, then it is likely tocause the contact in the thrust bearings, and eventually, themalfunction of the hydrodynamic bearing assembly. The thrust plate 103has a plurality of grooves 105 for generating the dynamic pressureformed thereon. The sleeve (not shown in the drawing) is arranged aroundthe shaft 101 covering the outer surface thereof. The sleeve rotatesrelative to and around the shaft 101 with predetermined gaptherebetween, thereby to generate the radial dynamic pressure. Therelative rotation in cooperation with the grooves 105 also generates thethrust dynamic pressure between the end surface and the thrust plate103. A hollow cylindrical constraint member 106 is fit into the hollowspace of the shaft 101 so that the outer surface of the constraintmember 106 contacts with the inner surface of the shaft 101. Also, theconstraint member 106 has a female screw formed on an inner surface. Thefemale screw receives a bolt 108 as the fastening means inserted via thethrough-hole 104 of the thrust plate 103. A washer 109 is engaged on thebolt 108. The constraint member 106 has parallel grooves 107 relative tothe axis formed on the outer surface thereof for directing the adhesive,of which details are illustrated in FIG. 30(a). As can be seen, thegroove 106 has an open end and a closed end terminating in the midportion. Although the constraint member 106 according to the embodimenthas three grooves 107, any number of grooves may be formed on theconstraint member 106.

Referring back to FIG. 29, in order to fasten the constraint member 106within the hollow space of the shaft 101, the constraint member 106 isfirstly inserted within the hollow space of the shaft 101 by means of ajig, and aligned in a predetermined position, then, the adhesive isdirected through the open end into the grooves 107.

Next, the bearing assembly so constructed of the shaft 101 and theconstraint member 106 is fastened with the thrust plate 103 bypositioning the thrust plate 103 on the shaft 101 so that the femalescrew of the thrust plate 103 and the bolt 108 are concentricallyaligned, by inserting the bolt with a washer 109 via the through-hole104 and the female screw, and by fastening the bolt with the femalescrew of the constraint member 106. This ensures both of the smoothenedsurfaces of the shaft 101 and the thrust plate 103 to closely contactwith each other, thereby for securing thereof with the high accuracy.What is fastened by the bolt 108 is the constraint member 106, which isbonded to the hollow inner surface of the shaft 101. Thus, fastening thebolt 108 causes the shearing force along the axis but no undesired forcein the radial direction. Therefore, this fastening mechanism receives nodeformation causing the bearing assembly to the malfunction.

It should be noted that while the adhesive is directed into the groove107 formed on the outer surface of the constraint member 106, the groove107 is not extending to and from both ends of the constraint member 106.In case where the groove 107 is extending to and from both ends of theconstraint member 106, there are a couple portions in which the adhesivemay unstick due to the external force such as the over-fastening of thebolt 108 or unexpected force to the shaft 101. The portions in which theadhesive may unstick include a portion between the adhesive and thegroove 107, a portion within the adhesive, and the portion between theadhesive and the shaft 101. However, the groove 107 having an open endand a closed end terminating in the mid portion as illustrated in thedrawing limits the portions in which the adhesive may unstick only tothe portion between the adhesive and the shaft 101. Thus, this fasteningmechanism reduces the risk of the adhesive to unstick. Alternately, thegrooves may be formed to have a spiral configuration, even though suchspiral configuration requires more tasks to achieve.

The experiments that the present inventors has conducted revealed thatafter assembling the fastening mechanism shown in FIG. 29, and fasteningthe bolt 108 with a torque of 100 kilograms centimeters has shifted theperpendicularity between the shaft 101 and the thrust plate 103 by 0.02microns, which falls within the tolerance limits thereof. Fastening thebolt 108 with a torque of 200 kilograms centimeters has caused theadhesive to unstick.

According to the embodiment, the constraint member 106 has the hollowspace for receiving the bolt 108 as the fastening member, however, theroles can be replaced to each other. FIG. 30(b) shows another constraintmember 106 a, which has a bolt-like configuration and a column-like headportion, on which one or more grooves 107 a are formed with the open endand the closed end terminating in the mid portion. In this instance, thefastening member is a nut (not shown) instead of the bolt. The nut andthe washer 109 engage on an outer screw of the constraint member 106 afor fastening each other.

FIG. 31 illustrates an alternative fastening mechanism using anotherconstraint member 116 extending along the axis. In the drawing, thethrust plate 113 has a relatively large through-hole 114 having adiameter same as the hollow space of the shaft 101. The constraintmember 116 extending along the axis to the shaft 101 is fit within thethrough-hole 114. The bolt in cooperation with a large washer 119 isfastened on the constraint member 116. The other mechanism and the wayto assemble are similar to those to the example shown in FIG. 29.

The advantage of the bearing assembly according to the embodiment isthat the shaft 101 and the thrust plate 103 can concentrically bearranged in an accurate manner because the constraint member 116 extendsto align itself with the thrust plate 103. The experiment that thepresent inventors have conducted revealed that the shaft 101 and thethrust plate 103 can concentrically be arranged with the accuracy of 10microns when the constraint member 116 has the extended portion with alength of 1 millimeter. Also, it should be noted that the constraintmember 116 may have the bolt-like configuration and the fastening membermay have the nut-like configuration to replace the roles of theconstraint member and the fastening member with each other.

Eleventh Embodiment

The eleventh embodiment of the hydrodynamic bearing assembly accordingto the present invention, which addresses improving the bearing rigidityby assuring the accuracy of the perpendicularity between the shaft andthe thrust plate, will be described hereinafter, with reference todrawings. FIG. 32 illustrates the hydrodynamic bearing assembly of theembodiment. In the drawing, a shoulder portion 115 extending towards thebearing axis is formed on the inner surface of the hollow shaft 111. Theother components are similar to those used for the bearing assembly ofthe previous tenth embodiment, and also, similar components are denotedas similar reference numerals.

The way to assemble the hydrodynamic bearing assembly of the embodimentwill be described herein. The constraint member 106 is inserted from thetop (the opposite end to the shoulder portion 115) into the shaft 111until it engages with the shoulder portion 115. According to theembodiment, the groove 107 formed on the outer surface of the constraintmember 106 also has the open end and the closed end, but openingupwardly rather than downwardly as the groove of the tenth embodiment.In other words, the open end extends in the opposite direction to theinserting direction. After the constraint member 106 is inserted toengage with the shoulder portion 115, the adhesive is directed from theupper end in the drawing, so that the constraint member 106 is bondedwithin the hollow space of the shaft 111. Then, the bolt 108 of thefastening member is put into the female screw of the constraint member106 and fastened with thrust plate 103 intervened between the washer 109and the shaft 111. The adhesive provided in the grooves operates asstops preventing the constraint member 106 when fastening the bolt 108.The fastening force is applied to the shaft 111 via the shoulder portion115 rather than via the adhesive. Thus, the groove 107 may extendsthrough the constraint member 106 from the top to the bottom thereof tohave both ends open, rather than one closed end terminating in the midportion.

Advantageously, the hydrodynamic bearing assembly of the embodiment hasa fastening strength between the shaft 111 and thrust plate 103, greaterthan that of the tenth embodiment because of the shoulder portion 115intervened therebetween. Since the shoulder portion 115 is locallyprovided adjacent the fastening portion, fastening the bolt 108 causesno elastic deformation of the shaft 111 along the axis. Also, fasteningthe bolt 108 causes a negligible extent of the radial expansion of theshaft 111. Thus, the bearing assembly has no adverse impact. To thisend, the hydrodynamic bearing assembly of the embodiment achieves theexcellent fastening strength.

The present inventors have conducted an experiment as follows. Thehydrodynamic bearing assembly was assembled as shown in FIG. 32, inwhich the constraint member 106 and the bolt 108 are fastened throughthe shoulder 115 in the hollow space by the torque of 100 kilogramscentimeters. Then, the perpendicularity was shifted by 0.02 microns,which falls within the acceptable range. No radial expansion of theouter surface of the shaft 111 was observed. Also, it should be notedthat the constraint member 116 may have the bolt-like configuration andthe fastening member may have the nut-like configuration to replace theroles of the constraint member and the fastening member with each other.

FIG. 33 shows an alternative example of the hydrodynamic bearingassembly of the embodiment. In the drawing, the shaft 121 has theshoulder portion 125 on the inner surface thereof. The constraint member116 is provided with an extension portion 127 extending through theshoulder portion 125 along the axis. The extension portion 127 protrudesbeyond the bottom surface of the shaft 121 for close fitting within thethrough-hole 124 of the thrust plate 123. In this embodiment, theconstraint member 126 and the fastening member are designed to have thebolt-like and nut-like configurations, respectively. Also, the malescrew 128 extrudes from the extension portion 127 of the constraintmember 126 via the through-hole 124 of the thrust plate 123, and engageswith the nut 118. However, as discussed above, the constraint member 126and the fastening member may be designed to have the nut-like andbolt-like configurations, respectively, so that roles of the constraintmember and the fastening member can be replaced with each other. Similarcomponents are denoted by similar reference numerals.

The way to assemble the hydrodynamic bearing assembly of the embodimentwill be described herein. The constraint member 126 is inserted from thetop (the opposite end to the shoulder portion 125) into the shaft 121until it engages with the shoulder portion 125. The groove 107 formed onthe outer surface of the constraint member 126 also has the upward openend. After the constraint member 126 is inserted to engage with theshoulder portion 125, the adhesive is directed from the upper end in thedrawing, so that the constraint member 126 is bonded within the hollowspace of the shaft 121. The extension portion 127 of the constraintmember 126 protrudes beyond the bottom surface of the shaft 121. Theouter surface of the extension portion 127 is processed with thesmoothness enough for close fitting within the through-hole 124 of thethrust plate 123, so that the extension portion 127 is closely fitwithin the through-hole 124 of the thrust plate 123. Also, the malescrew 128 extrudes from the extension portion 127 of the constraintmember 126 via the through-hole 124 of the thrust plate 123, and engageswith the nut of the fastening member 118 via the washer 109 forfastening the thrust plate 123 intervened therebetween.

Advantageously, the hydrodynamic bearing assembly of the embodiment hasthe greater fastening strength between the shaft 121 and thrust plate123, because of the shoulder portion 125 intervened therebetween. Also,advantageously, the hydrodynamic bearing assembly assures the accuracyof the perpendicularity between the shaft 121 and the thrust plate 123.This structure provides no adverse impact on the operation of thebearing assembly as the other aforementioned embodiments do.

Twelfth Embodiment

The twelfth embodiment of the hydrodynamic bearing assembly according tothe present invention, which addresses improving the bearing rigidity byassuring the accuracy of the perpendicularity between the shaft and thethrust plate, will be described hereinafter, with reference to drawings.FIG. 34 illustrates the hydrodynamic bearing assembly of the embodiment.In the drawing, a pair of fastening mechanisms is provided on the topand bottom ends of the shaft.

The hydrodynamic bearing assembly of FIG. 34 is constructed by a pair ofthe fastening mechanism described with reference of FIG. 31, using twosets of the thrust plates 113, the constraint members 116, the bolts108, and the washers 109. However, the present invention is not limitedto the combination of the fastening mechanisms. Thus, any combinationsof the fastening mechanisms, including ones according to the tenthand/or eleventh embodiments. However, if one of the fastening mechanismshas the shoulder portion 115 or 125, then the other one of the fasteningmechanisms cannot structurally have the shoulder portion 115 or 125.

A sleeve 102 is rotatably arranged around the outer surface of the shaft101. The shaft 101 is provided with a pair of thrust plates 113 on thetop and bottom surfaces thereof. There are predetermined gaps betweenthe shaft 101 and the sleeve 102 and between the thrust plates 113 andthe sleeve 102. After one of the thrust plates 113 is fastened onto theshaft 101, the sleeve 102 is arranged around the shaft 101, then theanother one of the thrust plate 113 is fastened onto the shaft 101. Atthis stage, the shaft 101 can be secured on a base as the stationarymember, and the sleeve 102 can be secured with the rotor 130 indicatedby the dashed lines as the rotational member. Alternatively, the shaft101 and the sleeve may be constructed as the rotational member and thestationary member, respectively. The hydrodynamic bearing assemblyconstructed with use of the fastening mechanisms as described aboveimproves the thrust rigidity. Further, the tilt rigidity can also beimproved by incorporating the pump-out type thrust bearing for one ofthe pair of thrust bearings.

According to each embodiment described above, the shaft is secureddirectly on the thrust plate. However, the shaft or the sleeve may beattached to an intervening member that is secured with the thrust plate.This invention can equally be applied to any type of the attachmentbetween the sleeve and the intervening member.

The fifth aspect according to the present invention is to improve theactivation feature of the hydrodynamic bearing assembly. As describedabove, the relative rotation with contact at the start of rotationcauses friction, heat, and energy consumption due to the increasedactivating torque.

Thirteenth Embodiment

The thirteenth embodiment of the hydrodynamic bearing assembly accordingto the present invention, which mainly addresses improving theactivation feature, will be described hereinafter, with reference todrawings. FIG. 35 is an enlarged view of the hydrodynamic bearingassembly incorporated with the spindle motor of FIG. 17. In FIG. 35, theshaft 131 is provided with the disk-shaped thrust plate 133 extendingperpendicularly to the axis of the shaft 131. The sleeve 132 is arrangedso as to surround the outer surface of the shaft 131 and the thrustplate 133 with predetermined gaps. A plurality of grooves for generatingthe dynamic pressure is formed either on the thrust plate 133 and theopposite surface of the sleeve 132 opposing to the thrust plate 133. Inother words, a thrust bearing is formed between a pair of the radialbearings, in which both bearings are communicated in fluid with eachother.

According to the embodiment, fillets 135 are provided adjacent to theconnecting portion between the shaft 131 and the thrust plate 133.Meanwhile, the sleeve 132 is provided with curved portions 136corresponding to the fillets 135. The fillets and the curved portionsmay be formed with arc configurations in the drawing, however, thepresent invention cannot be limited to the configuration. However,preferably, when the configuration is not arc, preferably the curve iscontinuous without bending points such that it causes no barrier of thefluid flow generating the dynamic pressure, and prevents the dust fromcovering thereon.

The drawing illustrates the shaft 131 and the sleeve 132 so that theradial gap between the shaft 131 and the sleeve 132 in both sides is thesame as each other. Similarly, the drawing illustrates the sleeve 132and the thrust plate 133 such that the thrust gap between the sleeve 132and the thrust plate 133 in both ends is the same as each other. Thus,each of the radial gap and the thrust gap is F/2 and D/2 when the totalradial gap and the total thrust gap are F and D, respectively. Also, theminimum distance from the fillet 135 to the curved portion 136 isdefined as S/2 in the drawing. According to the embodiment, the minimumdistance S/2, which is referred to as a fillet gap, satisfies thefollowing conditions;(F/2)<(S/2)<(D/2), thus,F<S<D

When starting to rotate the hydrodynamic bearing assembly,disadvantageously, the friction between the rotational member and thestationary member causes an electric power consumption and wears thethrust bearings. Preferably, the rotational member floats from thestationary member at lowest rotation rate. According to the conventionalbearing assemblies, in general, the thrust plate and thrust opposingsurface are in full contact with each other when halted, thus thefriction force is high. However, according to the embodiment, a part ofthe fillet 135 contacts with a part of the curved portion 136 asillustrated in FIG. 36 when halted if the following conditions are met;F<S<DThus, the full contact between the thrust plate and thrust opposingsurface can be avoided. Since the contact point between the fillet 135and the curved portion 136 is relatively close to the bearing axis, thearm length of the rotation moment is relatively short. Therefore, theactivating rotation force can be reduced in comparison with that of theconventional bearing assemblies having the full contact between thethrust plate 133 and thrust opposite 132 surface, and the rotationalmember can float at the earlier stage. In other words, the lessactivating rotation force causes the hydrodynamic bearing assembly to bemore compact and reduces the less power consumption. Further, the thrustbearing can wear less, and the disorder of the bearing assembly due tothe abrasion coom can be reduced. Even further, since the fillet portionmoves on the curved portion, which is smooth and continuous, thehydrodynamic bearing assembly of the embodiment cannot accumulate thedust as the conventional bearing assembly does on the edged corner.

If the fillet 135 and the curved portion 136 are too large, then thesufficient dynamic pressure cannot be obtained, thus they cannot belarger than a predetermined sizes in nature. The experiment that thepresent inventors has conducted revealed that, in the cross sectionincluding the bearing axis, the total curving length defined by thefillet 135 and the curved portion 136 is preferably two-third (⅔) orless of the total bearing length subtracting the total curved length,where the total bearing length is defined by the opposing surfacesbetween the shaft 131 and the sleeve 132 and between the sleeve 132 andthe thrust plate 133 in the radial and thrust bearings, respectively.

The fillet 135 and the curved portion 136 of FIG. 35 continuouslyconnecting between the radial and thrust bearings may be in forms ofcorn portions 135 a and 136 a as illustrated in FIG. 37. In FIG. 37, theshaft 131 and the sleeve 132 are positioned so that the radial gapsbetween the shaft 131 and the sleeve 132 in both sides are the same aseach other, and the sleeve 132 and the thrust plate 133 are positionedso that the thrust gaps between the sleeve 132 and the thrust plate 133in both ends are the same as each other. Thus, the radial gap and thethrust gap are F/2 and D/2. Also, the minimum distance between bothfrustum portions is defined as S/2. According to the embodiment, theminimum distance S/2, which is referred to as a corn gap, satisfies atany points the following conditions;(F/2)<(S/2)<(D/2), thus,F<S<D

When the above conditions are met, both of the corn portions partiallycontact with each other, and since the contact point therebetween isrelatively close to the bearing axis, the activating rotation force canbe reduced and the rotational member can float at the earlier stage.

In FIG. 37, while the intersection of the shaft 131 and the corn portion135 a and the intersection of the thrust plate 135 and the corn portion135 a are illustrated as being edge-shaped, those intersectionspreferably have minor round portions to promote the fluid flow and toavoid accumulation of the dust at the intersections. Also, similar tothe case of the curved portion, the total frustum length defined by thefrustum portions 136 a is preferably two-third (⅔) or less of the totalbearing length subtracting the total corn length.

The above-mentioned features are applied to the hydrodynamic bearingassembly of FIG. 38, which comprises a pair of the thrust bearings andthe radial bearing intervened therebetween. The hydrodynamic bearingassembly shown in FIG. 38 is constructed similarly to one shown in FIG.19 except that the continuous and smooth curved portions 135′ and 136′are formed at the connections of the radial bearing and the thrustbearing. Dimensions and reference numerals of the radial and thrustbearings of FIG. 38 are similar to those of FIG. 37. The gaps at theconnections, each of which defines a conduit of the fluid flow, have theminimum distance between the both sides (S/2). The activation featurecan be improved if the following conditions for (S/2) at any points aremet;(F/2)<(S/2)<(D/2), thus,F<S<DAs described above, the curved portions may be in form of the cornportions.

FIG. 39 illustrates the hydrodynamic bearing assembly with a generalstructure, which includes the radial bearing and only one thrust bearingformed at one end along the axis. The hydrodynamic bearing assemblycomprises a column shaft 137 having the outer surface parallel to theaxis, a hollow cylindrical sleeve 138 arranged around the outer surfaceof the shaft 137, a disk-shaped thrust plate 139 perpendicularly securedonto the shaft 137. In the general hydrodynamic bearing assembliesincluding such complex hydrodynamic bearing assembly, when the rotationis halted, preferably, the full contact between the thrust plate 139 andthrust opposing surface of the sleeve 138 is preferably avoided and theycontact at only at limited portion adjacent to the axis, so that theactivation feature is improved. In order to realize this, the cornerportion of the sleeve 138 and the root portion connecting the shaft 137and the thrust plate 139 at the intersections of the thrust bearing andthe radial bearing in FIG. 39, have the continuous and smooth curvedportions. When the rotation is halted, both of the curved portionscontact with each other. A first and second distances m and n can bedefined between two ascending positions of the curve portion of thesleeve 138 and the thrust plate 139, respectively. As describes above,the activation feature is improved, if the following condition is met;m<nIf the other type of the hydrodynamic bearing assemblies such as onesillustrated in FIGS. 35, 37, and 38 satisfy the above condition, thenthey also improve their activation features.

Although the configuration of each curved portion is not necessarily ina form of the arc, preferably, it has no bending point at least, and ithas a continuous curve or frustum portion for smoothly connecting thestraight shaft and the thrust plate. The curve or frustum portionadvantageously is likely to prevent accumulation of the dust at theintersection of thrust and radial bearings. Similar to the embodimentsdescribed above, the total curve or frustum length defined by the curveor frustum portions is preferably two-third (⅔) or less of the totalbearing length subtracting the total corn or frustum length.

The present inventors have conducted an experiment as follows. Thehydrodynamic bearing assembly shown in FIG. 35 was prepared includingthe thrust gap of 5 microns and the radial gap of 1 micron, alsoincluding the corner portion only of the sleeve 132 chamfered with theradius of 1 millimeter. Although no contacting noise of the bearingassembly was observed when the oscillation force by hand was applied tothe shaft, the dust was accumulated at the root portion defined betweenthe shaft 131 and the thrust plate 133. This revealed that the conditionof S>D, wherein S and D are illustrated in FIGS. 35 and 37,disadvantageously causes the dust accumulation.

Meanwhile, another hydrodynamic bearing assembly was prepared similarlyto the above except that it includes the root portion chamfered with theradius of 1 millimeter and the corner portion of the sleeve 132 with theperpendicular edge. Then, the oscillation force by hand was applied tothe shaft, the contacting noise of the bearing assembly was heard. Thisrevealed that the condition of F>S, wherein F and S are illustrated inFIGS. 35 and 37, disadvantageously causes the contact of the thrustbearing. In the embodiment, the shaft is described as the rotationalmember, the present invention can be applied to the bearing assemblywith the stationary shaft.

Fourteenth Embodiment

The fourteenth embodiment of the hydrodynamic bearing assembly accordingto the present invention, which addresses improving the activationfeature, will be described hereinafter, with reference to drawings. Ascan be seen from the drawing of the conventional hydrodynamic bearingassembly such as FIG. 42, the spindle motor has no stop preventing therotational member from pulling off upwardly or stationary member fromcoming out downwardly. Therefore, in case where no means for preventingthe pull-off is provided, the rotational member may be pulled off duringthe rotation. Also, in order to narrow the thrust gap between the thrustplate and the thrust opposing surface thereby increasing the dynamicpressure in the thrust bearing, preferably the thrust opposing surfaceis biased downwardly to the thrust plate during the rotation.

According to the prior arts, the rotor magnet is designed to have acenter offset to that of the electromagnet as illustrated in FIG. 40(a),in order to prevent the pull-off and for narrow the thrust gap in thethrust bearing. FIG. 40(a) is an enlarged view of the rotor magnet 208attached on the rotor 207 of the spindle motor shown in FIG. 42 and theelectromagnet 209 secured on the base plate 200 opposing to the rotormagnet 208. In FIG. 40(a), the electromagnet 209 includes the core sothat the attraction force is generated between the rotor magnet 208 andthe electromagnet 209. According to the prior arts, the rotor magnet 208is offset against the electromagnet 209 to generate the attraction forcein the direction indicated by the arrow 211 a, which has the horizontalcomponent 212 a and the vertical component 213 a. The vertical component213 a directing downwardly biases the rotational member of the spindlemotor such as rotor 207 to the stationary member. The downward biasprevents the rotational member from pulling off and keeps the thrust gapnarrow for increasing the thrust dynamic pressure in the thrust bearing.

Meanwhile, in case where the rotor magnet is designed to have a centeroffset to that of the electromagnet, since the horizontal component 212a of the attraction force between the rotor magnet 208 and theelectromagnet 209 is the force for driving the spindle motor, thedriving force is not greater than the attraction force between the rotormagnet 208 and the electromagnet 209. This requires the larger volume ofthe rotor magnet 208 and the electromagnet 209 for the strongerattraction force. This eventually causes the whole dimension and weightof the spindle motor to be greater and heavier. To be even worse, thevertical component 213 a always biases downwardly to increase the weightof the rotational member, thus to require higher rotation rate forfloating the rotational member when the spindle motor is activated.

Contrary to this, since the hydrodynamic bearing assembly used for thespindle motor according to the present invention as illustrated in FIG.23 are sandwiched by a pair of the thrust bearing, the rotational membercannot be pulled off. Further, the thrust bearings provided on the bothends in the axis direction require no external biasing force, and thehydrodynamic bearing assembly itself always keeps the gapsappropriately. Therefore, no particular means is required, for example,for offsetting the center of the electromagnet 209 relative to that ofthe rotor magnet 108, as shown in FIG. 40(a).

FIG. 40(b) is an enlarged view of a part of the spindle motor of theembodiment, illustrating the rotor magnet 58 attached on the rotor 57shown in FIG. 23 and the electromagnet 59 secured on the base 50opposing to the rotor magnet 58. The center of the rotor magnet 58 isaligned to that of the electromagnet 59 so that the attraction force canfully be utilized as the driving force. Thus, the activation feature isimproved by effectively utilizing the attraction force between the rotormagnet 58 and the electromagnet 59. In addition, no vertical componentof the attraction force biases the rotational member downwardly, thefriction force is reduced in the thrust bearing defined by the thrustplate 53 a and thrust opposing surface 54 a as shown in FIG. 23. Thisprevents the seizure therebetween and improves the endurance of thespindle motor.

FIG. 40(c) is an enlarged view of a part of the spindle motor of theembodiment, illustrating the rotor magnet 58 having the center offset inthe direction reverse to one of the prior arts. In the embodiment of thehydrodynamic bearing assembly according to the present invention asillustrated in FIG. 23, because the rotational member cannot be pulledoff, such offsetting can be implemented. This cause the attraction forcein the direction indicated by the arrow 211, which has the horizontalcomponent 212 c and the vertical component 213 c. The vertical component213 c directing upwardly supports the rotational member such as therotor magnet 58 and the rotor 57. Foe example, the distance between therotor magnet 58 and the electromagnet 59 is 5 millimeters and the centeroffset is about 0.5 millimeters, then the weight of the rotationalmember appears to be reduced by approximately 1 to 5 newtons. The weightreduction effects the rotational member to readily float when thespindle motor is activated, so that the floating rotation rate isreduced. Thus, the friction force between the thrust plate 53 a andthrust opposing surface 54 a in the thrust bearing of FIG. 23 isadvantageously further reduced.

The sixth aspect according to the present invention relates to thehydrodynamic bearing assembly having the reduced volume and weight, inparticular, relates to using ceramics as the material for the variouscomponents of the hydrodynamic bearing assembly. The ceramics is good atthe anti-abrasion, the endurance, and the rigidity, and also the weighthereof is light.

Fifteenth Embodiment

The fifteenth embodiment of the hydrodynamic bearing assembly accordingto the present invention, which addresses reducing the volume andweight, will be described hereinafter, with reference to drawings. FIG.41 is an enlarged view of the hydrodynamic bearing assembly incorporatedwith the spindle motor of FIG. 23. In FIG. 41, the hydrodynamic bearingassembly comprises a shaft 141, a sleeve 142 rotatably arrangedsurrounding the outer surface of the shaft 141, a first and seconddonut-like thrust plates 143 a and 143 b secured on each end of thesleeve 142. Either one of the first and second thrust plates 143 a and143 b has the pump-out type thrust grooves for generating the thrustdynamic pressure, and the other one of them has either one of thepump-out type and the pump-in type thrust grooves. It should be notedthat the thrust grooves may be formed on the thrust opposing surfaces ofthe sleeve.

The shaft 141 of the embodiment includes a stepped core member 146 and ahollow cylindrical outer member 147, because of the process condition ofthe ceramics. The core member and the outer member 147, which are madeof metal such as stainless and ceramics, respectively, are close fit toeach other by shrink fitting, cooling fitting or bonding with theadhesive. The sleeve 142 and both thrust plates 143 a and 143 b mayintegrally be formed of ceramics, and preferably so it is. Portionsrelatively rotating in the radial and thrust bearings, at least, may bemade of ceramics and other portions may be made of other materials suchas metal. Ceramics materials which may be used for the portions are, forexample, alumina, zirconia, silicon carbide, silicon nitride, sialon,and so on.

For example, an alumina-based ceramics has the Young's modulus withinthe range of approximately 300 to 400 giga-Pascals, which is aboutdouble to the steel, and the specific gravity of 3.9 which is about halfof the steel. Thus, briefly speaking, the alumina-based ceramicsprovides the rigidity (rigidity) double with half mass in comparisonwith steel, in addition, shows the good anti-abrasion. If the bearingassembly is made of ceramics, it can reduce its volume and weight andcan improve its anti-abrasion and endurance in comparison with stainlesssteel, for example. In the hydrodynamic bearing assembly, since therotational member contacts with the stationary member when halted andthen starts rotating, the friction and the seizure may be caused betweenthe members. Making those members of ceramics showing the goodanti-abrasion readily avoids such problems. The hydrodynamic bearingassembly used for the spindle motor incorporated in the HDD is requiredto be formed and assembled with high accuracy. Ceramics materials areless susceptible to the plastic deformation and the elastic deformationthan metal materials. Thus, usage of the ceramics materials reduces thedeformations in processing so as to provide a precise hydrodynamicbearing assembly.

Although this embodiment is described with reference to the hydrodynamicbearing assembly of FIG. 41, the present invention is not limited tothis embodiment, rather applicable to any type of the hydrodynamicbearing assemblies, for example, as illustrated in FIGS. 17, 20, and 34.Also, each of opposing surfaces of the hydrodynamic bearing assembly ispreferably made of ceramics materials, however, some of them may beformed of ceramics materials.

Sixteenth Embodiment

Each embodiment of the hydrodynamic bearing assemblies according to thepresent invention is described above in detail. The sixteenth embodimentaccording to the present invention relates to a spindle motor, and amemory device and bar code reader incorporating the spindle motor. Asabove, usage of the hydrodynamic bearing assemblies according to thepresent invention provides the spindle motor realizing the rotation athigh rate and heavy load in a stable manner, and the good enduranceagainst the external motion. Further, usage of the spindle motorprovides the memory equipment and the bar code reader realizing therotation at high rate and heavy load in a stable and reliable manner.

ADVANTAGES OF THE INVENTION

In the first embodiment of the hydrodynamic bearing assembly accordingto the present invention, the half-whirl can effectively be eliminatedwithout reducing the radial dynamic pressure. This eliminates thepossibility of the contact between the rotational member and thestationary member due to the reduction of the translation rigidity sothat the hydrodynamic bearing assembly realizing the stable rotationwithout the half-whirl can be provided.

In the second embodiment of the hydrodynamic bearing assembly accordingto the present invention, an usage of the high dynamic pressure mainlygenerated in the radial bearing prevents the contact between therotational member and the stationary member due to the external motionso that the robust hydrodynamic bearing assembly supporting itselfagainst the external motions can be provided.

In the third embodiment of the hydrodynamic bearing assembly accordingto the present invention, a peak of the dynamic pressure can be formedat the circumference of the thrust bearing so that the robusthydrodynamic bearing assembly supporting itself against the externalmotions can be provided.

In the fourth embodiment of the hydrodynamic bearing assembly accordingto the present invention, the shaft and thrust plate can be secured toeach other with a precise perpendicularity so that the bearing rigidityis improved and the undesired contact between members especially in thethrust bearing can be avoided.

In the fifth embodiment of the hydrodynamic bearing assembly accordingto the present invention, the activation torque at the beginning of theactivation can be reduced, or the activation force can effectively beused fo the rotation so that the floating rotation rate can quickly beobtained. This contributes the endurance and the low energy consumptionof the hydrodynamic bearing assembly.

In the sixth embodiment of the hydrodynamic bearing assembly accordingto the present invention, usage of the ceramics material can provide thecompact and lightweight hydrodynamic bearing assembly, which has highendurance and the precise dimensions.

In the sixth embodiment of the hydrodynamic bearing assembly accordingto the present invention, a reliable spindle motor, a memory device, andbar code reader can be provided, which operates at high rotation rate ina stable manner and includes the robust bearing rigidity and theimproved activation features.

1-45. (canceled)
 46. A hydrodynamic bearing assembly, comprising: aradial bearing including a column shaft having an outer surface parallelto an axis, and a hollow cylindrical sleeve having an inner surfacerotatably arranged around the outer surface of said shaft, said radialbearing for generating a radial dynamic pressure due to a relativerotation between said sleeve and said shaft; and a thrust bearingincluding a thrust plate formed or secured onto either one of said shaftand said sleeve, and a thrust opposing surface formed or secured ontothe other one of said shaft and said sleeve, said thrust bearing forgenerating a thrust dynamic pressure due to the relative rotationbetween said thrust plate and said thrust opposing surface; wherein saidthrust bearing is a pump-out type, and conducts the fluid in the thrustbearing in the direction from the axis to circumference, and wherein atleast one longitudinal groove is formed on either one of the outersurface of said shaft and the inner surface of said sleeve, thelongitudinal groove extending parallel or inclined to the axis betweenboth ends along the axis.
 47. The hydrodynamic bearing assemblyaccording to claim 46, wherein the longitudinal groove formed in theradial bearing is interrupted in the mid portion of the shaft along theaxis where the shaft has a cross section of a substantial circle. 48.The hydrodynamic bearing assembly according to claim 46, wherein thelongitudinal groove has a depth ratio defined by a depth thereof to adiameter of the surface is within the range from approximately 0.0001 toapproximately 0.005.